Fluid Mechanics, 9th Edition

FLUID
This International Student Edition is for use outside of the U.S.
Mechanics
NINTH EDITION
Frank M. White Henry Xue
Fluid Mechanics
Fluid Mechanics
Ninth Edition
Frank M. White
University of Rhode Island
Henry Xue
California State Polytechnic University
FLUID MECHANICS
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About the Authors
Frank M. White is Professor Emeritus of Mechanical and Ocean Engineering at the
University of Rhode Island. He is a native of Augusta, Georgia, and did his undergraduate studies at Georgia Tech, receiving a B.M.E. degree in 1954. Then he attended
the Massachusetts Institute of Technology for an S.M. degree in 1956, returning to
Georgia Tech to earn a Ph.D. degree in mechanical engineering in 1959. He began
teaching aerospace engineering at Georgia Tech in 1957 and moved to the University
of Rhode Island in 1964. He retired in January 1998.
At the University of Rhode Island, Frank became interested in oceanographic and
coastal flow problems, and in 1966 he helped found the first Department of Ocean
Engineering in the United States. His research interests have mainly been in viscous
flow and convection heat transfer. Known primarily as a teacher and writer, he received
the ASEE Westinghouse Teaching Excellence Award in addition to seven University
of Rhode Island teaching awards. His modest research accomplishments include some
80 technical papers and reports, the ASME Lewis F. Moody Research Award in 1973,
and the ASME Fluids Engineering Award in 1991. He is a Fellow of the ASME and
for 12 years served as editor-in-chief of the ASME Journal of Fluids Engineering. He
received a Distinguished Alumnus award from Georgia Tech in 1990 and was elected
to the Academy of Distinguished Georgia Tech Alumni in 1994.
In addition to the present text, he has written three undergraduate textbooks: Fluid
Mechanics; Heat Transfer; and Heat and Mass Transfer. He has served as a consulting editor of the McGraw-Hill Encyclopedia of Science and Technology from 1992
until 2006, and on the ASME Publications Committee until 2009. He continues to be
inspired by his late wife, Jeanne, and lives in Narragansett, Rhode Island, with his
dog Jack and his cat Kerry.
Henry Xue is Professor of Mechanical Engineering at California State Polytechnic
University. He received his B.S. degree from Jiangsu University in China, and his
M.S. and Ph.D. degrees from the University of Tokyo in Japan. Prior to joining
California State Polytechnic University in 2000, he was on the mechanical engineering faculty of National University of Singapore.
Henry has authored and coauthored many technical papers in computational fluid
mechanics and heat transfer in built environments, microscale gaseous flow modeling
and simulation using DSMC, and microcombustor and thermophotovoltaic energy
systems. He is a member of the American Society of Mechanical Engineers and the
American Society of Heating, Refrigeration and Air-Conditioning Engineers. He lives
with his wife, Sophia, in Irvine, California.
v
To Jeanne
Contents
Preface xi
2.9
2.10
Chapter 1
Introduction 2
1.1
1.2
1.3
1.4
1.5
1.6
1.7
1.8
1.9
1.10
1.11
Preliminary Remarks 3
The Concept of a Fluid 4
The Fluid as a Continuum 6
Dimensions and Units 7
System and Control Volume 16
Thermodynamic Properties of a Fluid 18
Viscosity and Other Secondary Properties 25
Flow Patterns: Streamlines, Pathlines, and Streaklines 41
Basic Flow Analysis Techniques 44
The Fundamentals of Engineering (FE) Examination 45
The History of Fluid Mechanics 46
Summary 46
Problems 47
Fundamentals of Engineering Exam Problems 55
Comprehensive Problems 55
References 58
Chapter 3
Integral Relations for a Control Volume 134
3.1
3.2
3.3
3.4
3.5
3.6
3.7
Chapter 2
Pressure Distribution in a Fluid 60
2.1
2.2
2.3
2.4
2.5
2.6
2.7
2.8
Pressure and Pressure Gradient 61
Equilibrium of a Fluid Element 63
Hydrostatic Pressure Distributions 64
Application to Manometry 71
Hydrostatic Forces on Plane Surfaces 75
Hydrostatic Forces on Curved Surfaces 82
Hydrostatic Forces in Layered Fluids 85
Buoyancy and Stability 88
Pressure Distribution in Rigid-Body Motion 93
Pressure Measurement 101
Summary 105
Problems 105
Word Problems 128
Fundamentals of Engineering Exam Problems 128
Comprehensive Problems 129
Design Projects 131
References 132
Basic Physical Laws of Fluid Mechanics 135
The Reynolds Transport Theorem 139
Conservation of Mass 147
The Linear Momentum Equation 152
Frictionless Flow: The Bernoulli Equation 168
The Angular Momentum Theorem 178
The Energy Equation 184
Summary 195
Problems 196
Word Problems 224
Fundamentals of Engineering Exam Problems 224
Comprehensive Problems 225
Design Project 227
References 227
Chapter 4
Differential Relations for Fluid Flow 228
4.1
4.2
The Acceleration Field of a Fluid 230
The Differential Equation of Mass Conservation
232
vii
viii
4.3
4.4
4.5
4.6
4.7
4.8
4.9
4.10
Contents
The Differential Equation of Linear Momentum 238
The Differential Equation of Angular Momentum 245
The Differential Equation of Energy 246
Boundary Conditions for the Basic Equations 249
The Stream Function 255
Vorticity and Irrotationality 262
Frictionless Irrotational Flows 264
Some Illustrative ­Incompressible Viscous Flows 270
Summary 279
Problems 279
Word Problems 290
Fundamentals of Engineering Exam Problems 291
Comprehensive Problems 291
References 292
Chapter 5
Dimensional Analysis and Similarity 294
5.1
5.2
5.3
5.4
5.5
Introduction 295
The Principle of Dimensional Homogeneity 299
The Pi Theorem 301
Nondimensionalization of the Basic Equations 312
Modeling and Similarity 321
Summary 333
Problems 334
Word Problems 342
Fundamentals of Engineering Exam Problems 342
Comprehensive Problems 343
Design Projects 344
References 345
Chapter 6
Viscous Flow in Ducts 346
6.1
6.2
6.3
6.4
6.5
6.6
6.7
6.8
6.9
Reynolds Number Regimes 347
Internal Viscous Flows 352
Head Loss—The Friction Factor 354
Laminar Fully Developed Pipe Flow 356
Turbulence Modeling 359
Turbulent Pipe Flow 366
Four Types of Pipe Flow Problems 374
Flow in Noncircular Ducts 380
Minor or Local Losses in Pipe Systems 389
6.10
6.11
6.12
Multiple-Pipe Systems 398
Experimental Duct Flows: Diffuser Performance 404
Fluid Meters 409
Summary 431
Problems 432
Word Problems 451
Fundamentals of Engineering Exam Problems 451
Comprehensive Problems 452
Design Projects 454
References 455
Chapter 7
Flow Past Immersed Bodies 458
7.1
7.2
7.3
7.4
7.5
7.6
7.7
Reynolds Number and ­Geometry Effects 459
Momentum Integral Estimates 463
The Boundary Layer ­Equations 467
The Flat-Plate Boundary Layer 469
Boundary Layers with Pressure Gradient 479
Drag of Two- and Three-Dimensional Bodies 485
Forces on Lifting Bodies 504
Summary 513
Problems 514
Word Problems 527
Fundamentals of Engineering Exam Problems 527
Comprehensive Problems 528
Design Project 529
References 529
Chapter 8
Potential Flow 532
8.1
8.2
8.3
8.4
8.5
8.6
8.7
8.8
Introduction and Review 533
Elementary Plane Flow Solutions 536
Superposition of Plane Flow Solutions 544
Plane Flow Past Closed-Body Shapes 550
Other Plane Potential Flows 559
Images 563
Airfoil Theory 566
Axisymmetric Potential Flow 574
Summary 580
Problems 580
Word Problems 590
Contents ix
Comprehensive Problems 590
Design Projects 591
References 591
Chapter 9
Compressible Flow 594
9.1
9.2
9.3
9.4
9.5
9.6
9.7
9.8
9.9
9.10
Introduction: Review of Thermodynamics 596
The Speed of Sound 600
Adiabatic and Isentropic Steady Flow 603
Isentropic Flow with Area Changes 609
The Normal Shock Wave 616
Operation of Converging and Diverging Nozzles 624
Compressible Duct Flow with Friction 629
Frictionless Duct Flow with Heat Transfer 640
Mach Waves and Oblique Shock Waves 645
Prandtl–Meyer Expansion Waves 655
Summary 668
Problems 669
Word Problems 682
Fundamentals of Engineering Exam Problems 682
Comprehensive Problems 683
Design Projects 684
References 685
Chapter 10
Open-Channel Flow 686
10.1
10.2
10.3
10.4
10.5
10.6
10.7
Introduction 687
Uniform Flow; The Cheˊzy Formula and
the Manning Formula 693
Efficient Uniform-Flow Channels 699
Specific Energy; Critical Depth 702
The Hydraulic Jump 710
Gradually Varied Flow 714
Flow Measurement and Control by Weirs 722
Summary 730
Problems 730
Word Problems 742
Fundamentals of Engineering Exam Problems
Comprehensive Problems 743
Design Projects 744
References 744
743
Chapter 11
Turbomachinery 746
11.1
11.2
11.3
11.4
11.5
11.6
Introduction and Classification 747
The Centrifugal Pump 750
Pump Performance Curves and Similarity Rules 756
Mixed- and Axial-Flow Pumps: The Specific Speed 767
Matching Pumps to System Characteristics 775
Turbines 782
Summary 796
Problems 797
Word Problems 810
Comprehensive Problems 810
Design Project 812
References 812
Appendix A Physical Properties of Fluids 814
Appendix B Compressible Flow Tables 819
Appendix C Conversion Factors 826
Appendix D Equations of Motion in Cylindrical Coordinates 828
Appendix E Estimating Uncertainty in Experimental Data 830
Appendix F Numerical Methods 832
Answers to Selected Problems 846
Index 853
Conversion Factors 864
Moody Chart 866
Preface
General Approach
The book is intended for an undergraduate engineering course in fluid mechanics. The
principles considered in the book are fundamental and have been well established in
the community of fluids engineering. However, in presenting this important subject,
we have drawn on our own ideas and experience. There is plenty of material for a
full year of instruction, and the content can also easily be divided into two semesters
of teaching. There have been some additions and deletions in this ninth edition of
Fluid Mechanics, but no philosophical change. There are still eleven chapters, plus
appendices. The informal, student-oriented style is retained and, if it succeeds, has
the flavor of an interactive lecture by the authors.
New co-author Dr. Henry Xue was brought on board for this edition.
Learning Tools
The total number of problem exercises continues to increase, from 1089 in the first
edition, to 1681 in the ninth edition. Most of these are basic end-of-chapter problems,
sorted according to topic. There are also Word Problems, multiple-choice Fundamentals of Engineering Problems, Comprehensive Problems, and Design Projects. Answers
to Selected Problems, at the end of the book, provides the answers to approximately
700 end-of-chapter problems.
In addition, there are many example problems throughout the chapters that showcase the recommended sequence of problem-solving steps outlined in Section 1.7.
Most of the problems in this text can be solved with a hand calculator. Some can
even be simply explained in words. A few problems, especially in Chapters 6, 7, 9,
and 10, involve solving complicated algebraic expressions, that would be laborious
for hand calculation but can be much more easily handled using licensed equationsolving software. The authors have provided examples of how to solve complicated
example problems using Microsoft Excel, as illustrated in Example 6.5. Excel contains
several hundred special mathematical functions for performing engineering and statistics calculations.
Content Changes
The overall content and order of presentation have not changed substantially in this
edition except for the following:
xi
xii
Preface
Chapter 1 renames Section 1.5 “System and Control Volume.” Definitions of system
and control volume, which formerly were scattered over many chapters, are now consolidated in this section. A new subheading, “Methods of Description,” has been added.
The Lagrangian and Eulerian methods of description have been moved here from Chapter 4. Discussions of velocity and acceleration fields are retained as examples of using
the control volume approach with the Eulerian method of description. The section
“Flow Patterns: Streamlines, Streaklines, and Pathlines,” formerly Section 1.9, has been
moved forward as Section 1.8 for better continuity in the introduction of fluid and flow
systems. A new subsection, “Integral and Differential Approaches,” has been added to
the new Section 1.9, “Basic Flow Analysis Techniques.”
Chapter 2 edits descriptions in Section 2.4, “Application to Mamometry,” using the
methods of “pressure increasing downward” and “jump across” typically. The coordinates for Figure 2.2 have been reset to be consistent with Figure 2.1. Figure 2.12 has
been replaced with a new figure to better illustrate the pressure distribution on a
submerged surface.
Chapter 3 adds three subheadings to elaborate areas where the linear momentum
equation can be applied. Example 3.7 has been rewritten to better demonstrate how to
solve the anchoring forces on a piping elbow. Brief discussions have been added to
examples of the sluice gate and impinging jet with relative velocity for an inertial, moving, and nondeforming control volume.
Chapter 4 adds the constant heat flux boundary condition to the energy equation.
Inlet and outlet boundary conditions are separated because the free-flow conditions are
more common at the outlet. New Example 4.10 investigates the rotation of a Couette
flow and a “potential vortex” flow.
Chapter 5 carries the topics of Section 5.2—the choice of variables and scaling
parameters—into Section 5.3 to make it easier for students to follow the arguments.
The topic “Some Peculiar Engineering Equations” has been removed from Section 5.2
because most of those equations will be introduced in Chapter 10.
In Chapter 6, Section 6.2 has been retitled “Internal Viscous Flow.” Brief discussions
have been added to four types of pipe flow problems to guide students in applying
appropriate strategies for designing pipe systems.
In Chapter 7, the discussion in the section “Transition to Turbulence” in Section 7.4
has been improved. The classification of external flow is elaborated. Former Section
7.6 has been split into two sections: “Drag” and “Forces on Lifting Bodies.” The methodology for solving an external flow problem is summarized.
An entire section of Chapter 8 on numerical methods, including problem exercises,
has been moved to new Appendix F. The vast majority of universities do not cover
numerical methods in a fundamental fluid mechanics course. Because the CFD methods
are becoming a powerful tool for solving almost all problems of fluid flow, it was also
inappropriate to place that topic at the end of this chapter. A new example of a free
vortex has been added to Section 8.2.
Chapter 9 clarifies why we can simplify compressible flow as one-dimensional
isentropic flow. Section 9.3 explains the identity of the momentum equation and the
energy equation for isentropic flows. Discussions have been added regarding how to
use the variables of stagnation pressure, density, and throat area after the shock wave
in calculation.
Preface xiii
Chapter 10 improves the physical interpretation of the Froude number in Section
10.1. There is a new subsection “Effects of Froude Number.” The need to maximize
the hydraulic radius in order to achieve an efficient channel is elaborated in Section
10.3.
Chapter 11 elaborates further on pump performance curves. New Figure 11.18a
illustrates the derivation for the system head. The data for worldwide wind power capacity have been updated.
Appendices A to E remain unchanged. The new Appendix F, “Numerical Methods,”
presents text that formerly was in Chapter 8. This will continue to serve instructors who
use this material for introducing the CFD methods to their students.
Additionally, this title is supported by SmartBook, a feature of the LearnSmart
adaptive learning system that assesses student understanding of course content through
a series of adaptive questions. This platform has provided feedback from thousands
of students, identifying those specific portions of the text that have resulted in the
greatest conceptual difficulty and comprehension among students. For the ninth edition, the entire text was reviewed and revised based on this LearnSmart student data.
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Acknowledgments
We wish to express our appreciation to the many people who have helped us in recent
revisions. Material help, in the form of photos, articles, and problems, came from Scott
Larwood of the University of the Pacific; Sukanta Dash of the Indian Institute of Technology at Kharagpur; Mark Coffey of the Colorado School of Mines; Mac Stevens of
Oregon State University; Stephen Carrington of Malvern Instruments; Carla Cioffi of
NASA; Lisa Lee and ­Robert Pacquette of the Rhode Island Department of Environmental Management; Vanessa Blakeley and Samuel Schweighart of Terrafugia Inc.; Beric
Skews of the University of the Witwatersrand, South Africa; Kelly Irene Knorr and
John Merrill of the School of Oceanography at the University of Rhode Island; Adam
Rein of Altaeros Energies Inc.; Dasari Abhinav of Anna University, India; Kris Allen
of Transcanada Corporation; Bruce Findlayson of the University of Washington; Wendy
Koch of USA Today; Liz Boardman of the South County Independent; Beth Darchi
and Colin McAteer of the American Society of Mechanical Engineers; Catherine Hines
of the William Beebe Web Site; Laura Garrison of York College of Pennsylvania.
Many others have supported us, throughout our revisions efforts, with comments
and suggestions: Barry Satvat of Northeastern University; Sangjin Ryu of University
of Nebraska–Lincoln; Edgar Caraballo of Miami University; Nigel Kaye of Clemson
University; Margaret Lang of Humboldt State University; Jie Xu of the University of
Illinois at Chicago; Joseph A. Schaefer of Iowa State University; Saeil Jeon of North
Carolina A&T State University; Diane DiMassa of Massachusetts Maritime Academy;
Angela Shih, Paul Nissenson, and Soorgul Wardak of California State Polytechnic
University.
The McGraw-Hill staff was, as usual, very helpful. We are indebted to Heather
Ervolino, Beth Bettcher, Shannon O’Donnell, and Jane Mohr.
Finally, special thanks go to our families for the continuing support. Frank is
especially grateful to Jeanne, who remains in his heart, and his sister Sally White
GNSH, his dog Jack, and his cat Kerry. Henry appreciates his wife Sophia for her
understanding with all the days that went into this effort.
Frank M. White
Henry Xue
xiv
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Falls on the Nesowadnehunk Stream in Baxter State Park, Maine, which is the northern
­terminus of the Appalachian Trail. Such flows, open to the atmosphere, are driven simply
by gravity and do not depend much upon fluid properties such as density and viscosity.
They are discussed later in Chap. 10. To the writer, one of the joys of fluid mechanics is
that visualization of a fluid flow process is simple and beautiful [Robert Cable/Natural
Selection/Design Pics].
2
Chapter 1
Introduction
1.1 Preliminary Remarks
Fluid mechanics is the study of fluids either in motion (fluid dynamics) or at rest
(fluid statics). Both gases and liquids are classified as fluids, and the number of
fluid engineering applications is enormous: breathing, blood flow, swimming,
pumps, fans, turbines, airplanes, ships, rivers, windmills, pipes, missiles, icebergs,
engines, filters, jets, and sprinklers, to name a few. When you think about it,
almost everything on this planet either is a fluid or moves within or near a fluid.
The essence of the subject of fluid flow is a judicious compromise between
theory and experiment. Since fluid flow is a branch of mechanics, it satisfies a
set of well-documented basic laws, and thus a great deal of theoretical treatment
is available. However, the theory is often frustrating because it applies mainly to
idealized situations, which may be invalid in practical problems. The two major
obstacles to a workable theory are geometry and viscosity. The basic equations
of fluid motion (Chap. 4) are too difficult to enable the analyst to attack arbitrary
geometric configurations. Thus most textbooks ­concentrate on flat plates, circular
pipes, and other easy geometries. It is possible to apply ­numerical computer
techniques to complex geometries, and specialized textbooks are now a­vailable
to explain the new computational fluid dynamics (CFD) ­approximations and
methods [1–4].1 This book will present many t­heoretical results while keeping
their limitations in mind.
The second obstacle to a workable theory is the action of viscosity, which can
be neglected only in certain idealized flows (Chap. 8). First, viscosity increases
the difficulty of the basic equations, although the boundary-layer approximation
found by Ludwig Prandtl in 1904 (Chap. 7) has greatly simplified viscous-flow
analyses. Second, viscosity has a destabilizing effect on all fluids, giving rise, at
frustratingly small velocities, to a disorderly, random phenomenon called turbulence. The theory of turbulent flow is crude and heavily backed up by experiment
(Chap. 6), yet it can be quite serviceable as an engineering estimate. This textbook
only introduces the standard experimental correlations for turbulent time-mean
flow. Meanwhile, there are advanced texts on both time-mean turbulence and
1
Numbered references appear at the end of each chapter.
3
4
Chapter 1 Introduction
turbulence modeling [5, 6] and on the newer, computer-intensive direct numerical
simulation (DNS) of fluctuating turbulence [7, 8].
Thus there is theory available for fluid flow problems, but in all cases it should
be backed up by experiment. Often the experimental data provide the main source
of information about specific flows, such as the drag and lift of immersed bodies
(Chap. 7). Fortunately, fluid mechanics is a highly visual subject, with good
instrumentation [9–11], and the use of dimensional analysis and modeling concepts (Chap. 5) is widespread. Thus experimentation provides a natural and easy
complement to the theory. You should keep in mind that theory and experiment
should go hand in hand in all studies of fluid mechanics.
1.2 The Concept of a Fluid
From the point of view of fluid mechanics, all matter consists of only two states,
fluid and solid. The difference between the two is perfectly obvious to the layperson, and it is an interesting exercise to ask a layperson to put this difference into
words. The technical distinction lies with the reaction of the two to an applied
shear or tangential stress. A solid can resist a shear stress by a static deflection;
a fluid cannot. Any shear stress applied to a fluid, no matter how small, will result
in motion of that fluid. The fluid moves and deforms continuously as long as the
shear stress is applied. As a corollary, we can say that a fluid at rest must be in
a state of zero shear stress, a state often called the hydrostatic stress condition in
structural analysis. In this condition, Mohr’s circle for stress reduces to a point,
and there is no shear stress on any plane cut through the element under stress.
Given this definition of a fluid, every layperson also knows that there are two
classes of fluids, liquids and gases. Again the distinction is a technical one concerning the effect of cohesive forces. A liquid, being composed of relatively
close-packed molecules with strong cohesive forces, tends to retain its volume
and will form a free surface in a gravitational field if unconfined from above.
Free-surface flows are dominated by gravitational effects and are studied in
Chaps. 5 and 10. Since gas molecules are widely spaced with negligible cohesive
forces, a gas is free to expand until it encounters confining walls. A gas has no
definite volume, and when left to itself without confinement, a gas forms an
atmosphere that is essentially hydrostatic. The hydrostatic behavior of liquids and
gases is taken up in Chap. 2. Gases cannot form a free surface, and thus gas flows
are rarely concerned with gravitational effects other than buoyancy.
Figure 1.1 illustrates a solid block resting on a rigid plane and stressed by its own
weight. The solid sags into a static deflection, shown as a highly exaggerated dashed
line, resisting shear without flow. A free-body diagram of element A on the side of
the block shows that there is shear in the block along a plane cut at an angle θ
through A. Since the block sides are unsupported, element A has zero stress on the
left and right sides and compression stress σ = −p on the top and bottom. Mohr’s
circle does not reduce to a point, and there is nonzero shear stress in the block.
By contrast, the liquid and gas at rest in Fig. 1.1 require the supporting walls
in order to eliminate shear stress. The walls exert a compression stress of −p and
reduce Mohr’s circle to a point with zero shear everywhere—that is, the hydrostatic ­condition. The liquid retains its volume and forms a free surface in the
1.2 The Concept of a Fluid 5
Free
surface
Static
deflection
A
A
Solid
A
Gas
Liquid
(a)
(c)
– σ1
θ
0
θ
τ1
τ=0
p
0
A
–σ = p
τ
τ
(1)
–p
σ
(b)
p
A
–σ = p
2θ
p
Hydrostatic
condition
σ
–p
(d )
Fig. 1.1 A solid at rest can resist shear. (a) Static deflection of the solid; (b) equilibrium
and Mohr’s circle for solid element A. A fluid cannot resist shear. (c) Containing walls are
needed; (d ) equilibrium and Mohr’s circle for fluid ­element A.
container. If the walls are removed, shear develops in the liquid and a big splash
results. If the ­container is tilted, shear again develops, waves form, and the free
surface seeks a horizontal ­configuration, pouring out over the lip if necessary.
Meanwhile, the gas is unrestrained and expands out of the container, filling all
available space. Element A in the gas is also hydrostatic and exerts a compression
stress −p on the walls.
In the previous discussion, clear decisions could be made about solids, liquids,
and gases. Most engineering fluid mechanics problems deal with these clear
cases—that is, the common liquids, such as water, oil, mercury, gasoline, and
alcohol, and the common gases, such as air, helium, hydrogen, and steam, in their
common temperature and pressure ranges. There are many borderline cases, however, of which you should be aware. Some apparently “solid” substances such as
asphalt and lead resist shear stress for short periods but actually deform slowly
and exhibit definite fluid behavior over long periods. Other substances, notably
colloid and slurry mixtures, resist small shear stresses but “yield” at large stress
6
Chapter 1 Introduction
and begin to flow as fluids do. Specialized textbooks are devoted to this study
of more general deformation and flow, a field called rheology [16]. Also, liquids
and gases can coexist in two-phase mixtures, such as steam–water mixtures or
water with entrapped air bubbles. Specialized textbooks present the analysis of
such multiphase flows [17]. Finally, in some situations the distinction between a
liquid and a gas blurs. This is the case at temperatures and pressures above the
so-called critical point of a substance, where only a single phase exists, primarily resembling a gas. As pressure increases far above the critical point, the gaslike
substance becomes so dense that there is some resemblance to a liquid, and the
usual thermodynamic approximations like the perfect-gas law become inaccurate.
The critical temperature and pressure of water are Tc = 647 K and pc = 219 atm
­(atmosphere)2 so that typical problems involving water and steam are below the
critical point. Air, being a mixture of gases, has no distinct critical point, but its
principal component, nitrogen, has Tc = 126 K and pc = 34 atm. Thus typical
problems involving air are in the range of high temperature and low pressure
where air is distinctly and definitely a gas. This text will be concerned solely
with clearly identifiable liquids and gases, and the borderline cases just discussed
will be beyond our scope.
1.3 The Fluid as a Continuum
We have already used technical terms such as fluid pressure and density without
a rigorous discussion of their definition. As far as we know, fluids are aggregations of molecules, widely spaced for a gas, closely spaced for a liquid. The
distance between molecules is very large compared with the molecular diameter.
The molecules are not fixed in a lattice but move about freely relative to each
other. Thus fluid density, or mass per unit volume, has no precise meaning
because the number of molecules occupying a given volume continually changes.
This effect becomes unimportant if the unit volume is large compared with, say,
the cube of the molecular spacing, when the number of molecules within the
volume will remain nearly constant in spite of the enormous interchange of particles across the boundaries. If, however, the chosen unit volume is too large,
there could be a noticeable variation in the bulk aggregation of the particles. This
situation is illustrated in Fig. 1.2, where the “density” as calculated from molecular mass δm within a given volume δ 𝒱 is plotted versus the size of the unit
volume. There is a limiting volume δ 𝒱* below which molecular variations may
be important and above which aggregate variations may be important. The density
ρ of a fluid is best defined as
ρ=
lim
δ 9→δ 9*
δm
δ9
(1.1)
The limiting volume δ 𝒱* is about 10−9 mm3 for all liquids and for gases at
atmospheric pressure. For example, 10−9 mm3 of air at standard conditions contains approximately 3 × 107 molecules, which is sufficient to define a nearly
constant density according to Eq. (1.1). Most engineering problems are concerned
2
One atmosphere equals 2116 lbf/ft2 = 101,300 Pa.
1.4 Dimensions and Units 7
ρ
Elemental
volume
ρ = 1000 kg/m3
ρ = 1200
Fig. 1.2 The limit definition of
­continuum fluid density: (a) an
­elemental volume in a fluid region
of variable continuum density;
(b) calculated density versus size
of the elemental volume.
Macroscopic
uncertainty
ρ = 1100
δυ
ρ = 1300
Microscopic
uncertainty
1200
0
δ𝒱* ≈ 10–9 mm3
δ𝒱
Region containing fluid
(a)
(b)
with physical dimensions much larger than this limiting volume, so that density
is essentially a point function and fluid properties can be thought of as varying
continually in space, as sketched in Fig. 1.2a. Such a fluid is called a continuum,
which simply means that its variation in properties is so smooth that differential
calculus can be used to analyze the substance. We shall assume that continuum
calculus is valid for all the analyses in this book. Again there are two borderline
cases for gases. One is at such low pressures that molecular spacing and mean
free path3 are comparable to, or larger than, the physical size of the system.
Applications include vacuum engineering, aero-thermal analysis and design of
spacecrafts, satellites, missiles, etc., flying at high altitudes. The non-continuum
effects also become significant when system length scales reduce to microscopically small. Applications with microscopic length scales are becoming increasingly common since the advent of Micro-Electro-Mechanical Systems (MEMS)
and nano devices, where the characteristic length of the system decreases to a
magnitude of sub-micron or nanometer. Both cases require that the continuum
approximation be dropped in favor of a molecular theory of rarefied gas flow
[18]. In principle, all fluid mechanics problems can be attacked from the molecular viewpoint, but no such attempt will be made here. Note that the use of
continuum calculus does not preclude the possibility of discontinuous jumps in
fluid properties across a free surface or fluid interface or across a shock wave in
a compressible fluid (Chap. 9). Our calculus in analyzing fluid flow must be
flexible enough to handle discontinuous boundary conditions.
1.4 Dimensions and Units
A dimension is the measure by which a physical variable is expressed quantitatively. A unit is a particular way of attaching a number to the quantitative dimension. Thus length is a dimension associated with such variables as distance,
displacement, width, deflection, and height, while centimeters and inches are both
numerical units for expressing length. Dimension is a powerful concept about
which a splendid tool called dimensional analysis has been developed (Chap. 5),
while units are the numerical quantity that the customer wants as the final answer.
3
The mean distance traveled by molecules between collisions (see Prob. P1.5).
8
Chapter 1 Introduction
Table 1.1 Primary Dimensions in SI and BG Systems
Primary dimension
Mass {M}
Length {L}
Time {T}
Temperature {Θ}
SI unit
BG unit
Kilogram (kg)
Meter (m)
Second (s)
Kelvin (K)
Slug
Foot (ft)
Second (s)
Rankine (°R)
Conversion factor
1
1
1
1
slug = 14.5939 kg
ft = 0.3048 m
s=1s
K = 1.8°R
In 1872 an international meeting in France proposed a treaty called the Metric
Convention, which was signed in 1875 by 17 countries including the United
States. It was an improvement over British systems because its use of base 10 is
the foundation of our number system, learned from childhood by all. Problems
still remained because even the metric countries differed in their use of kiloponds
instead of dynes or newtons, kilograms instead of grams, or calories instead of
joules. To standardize the metric system, a General Conference of Weights and
Measures, attended in 1960 by 40 countries, proposed the International System
of Units (SI). We are now undergoing a painful period of transition to SI, an
adjustment that may take many more years to complete. The professional societies have led the way. Since July 1, 1974, SI units have been required by all papers
published by the American Society of Mechanical Engineers, and there is a textbook explaining the SI [19]. The present text will use SI units together with
British gravitational (BG) units.
Primary Dimensions
In fluid mechanics there are only four primary dimensions from which all other
dimensions can be derived: mass, length, time, and temperature.4 These dimensions and their units in both systems are given in Table 1.1. Note that the Kelvin
unit uses no degree symbol. The braces around a symbol like {M} mean “the
dimension” of mass. All other variables in fluid mechanics can be expressed in
terms of {M}, {L}, {T}, and {Θ}. For example, acceleration has the dimensions
{LT −2}. The most crucial of these secondary dimensions is force, which is directly
related to mass, length, and time by Newton’s second law. Force equals the time
rate of change of momentum or, for constant mass,
F = ma
(1.2)
From this we see that, dimensionally, {F} = {MLT −2}.
The International System (SI)
The use of a constant of proportionality in Newton’s law, Eq. (1.2), is avoided
by defining the force unit exactly in terms of the basic units. In the SI system,
the basic units are kilograms {M}, meters {L}, and seconds {T}. We define
1 newton of force = 1 N = 1 kg · 1 m/s2
4
If electromagnetic effects are important, a fifth primary dimension must be included, electric
current {I}, whose SI unit is the ampere (A).
1.4 Dimensions and Units 9
The newton is a relatively small force, about the weight of an apple (0.225 lbf ).
In addition, the basic unit of temperature {Θ} in the SI system is the degree
Kelvin, K. They are referred to as the MLTΘ system of dimension. Use of these
SI units (kg, m, s, K) will require no conversion factors in our equations.
The British Gravitational (BG) System
In the BG system also, a constant of proportionality in Eq. (1.2) is avoided by
defining the force unit exactly in terms of the basic units. In the BG system, the
basic units are pound-force {F}, feet {L}, and seconds {T}. We define
1 pound of force = 1 lbf = 1 slug · 1 ft/s2
One lbf ≈ 4.4482 N and approximates the weight of four apples. We will use
the abbreviation lbf for pound-force and lbm for pound mass. The slug is a
rather hefty mass, equal to 32.174 lbm. The basic unit of temperature {Θ} in
the BG system is the degree Rankine, °R. Recall that a temperature difference
1 K = 1.8°R. They are referred to as the FLTΘ system of dimension. Use of
these BG units (lbf, ft, s, °R) will require no conversion factors in our equations.
Other Unit Systems
There are other unit systems still in use. At least one needs no proportionality
­constant: the CGS system (dyne, gram, cm, s, K). However, CGS units are too
small for most applications (1 dyne = 10−5 N) and will not be used here.
In the USA, some still use the English Engineering system (lbf, lbm, ft, s, °R),
where the basic mass unit is the pound of mass. Newton’s law (1.2) must be rewritten:
F=
ma
ft · lbm
, where gc = 32.174
gc
lbf · s2
(1.3)
The constant of proportionality, gc, has both dimensions and a numerical value
not equal to 1.0. The present text uses only the SI and BG systems and will not
solve problems or examples in the English Engineering system. Because Americans still use them, a few problems in the text will be stated in truly awkward
units: acres, gallons, ounces, or miles. Your assignment will be to convert these
and solve in the SI or BG systems.
The Principle of Dimensional Homogeneity
In engineering and science, all equations must be dimensionally homogeneous,
that is, each additive term in an equation must have the same dimensions. For
example, take Bernoulli’s incompressible equation, to be studied and used
throughout this text:
p+
1 2
ρV + ρgZ = constant
2
Each and every term in this equation must have dimensions of pressure {ML−1T −2}.
We will examine the dimensional homogeneity of this equation in detail in Example 1.3.
10
Chapter 1 Introduction
Table 1.2 Secondary Dimensions in Fluid Mechanics
Secondary dimension
Area {L2}
Volume {L3}
Velocity {LT −1}
Acceleration {LT −2}
Pressure or stress {ML−1T −2}
Angular velocity {T −1}
Energy, heat, work {ML2T −2}
Power {ML2T −3}
Density {ML−3}
Viscosity {ML−1T −1}
Specific heat {L2T −2Θ−1}
SI unit
BG unit
m2
m3
m/s
m/s2
Pa = N/m2
s−1
J=N·m
W = J/s
kg/m3
kg/(m · s)
m2/(s2 · K)
ft2
ft3
ft/s
ft/s2
lbf/ft2
s−1
ft · lbf
ft · lbf/s
slugs/ft3
slugs/(ft · s)
ft2/(s2 · °R)
Conversion factor
1
1
1
1
1
1
1
1
1
1
1
m2 = 10.764 ft2
m3 = 35.315 ft3
ft/s = 0.3048 m/s
ft/s2 = 0.3048 m/s2
lbf/ft2 = 47.88 Pa
s−1 = 1 s−1
ft · lbf = 1.3558 J
ft · lbf/s = 1.3558 W
slug/ft3 = 515.4 kg/m3
slug/(ft · s) = 47.88 kg/(m · s)
m2/(s2 · K) = 5.980 ft2/(s2 · °R)
A list of some important secondary variables in fluid mechanics, with dimensions derived as combinations of the four primary dimensions, is given in
Table 1.2. A more complete list of conversion factors is given in App. C.
EXAMPLE 1.1
A body weighs 1000 lbf when exposed to a standard earth gravity g = 32.174 ft/s2.
(a) What is its mass in kg? (b) What will the weight of this body be in N if it is
exposed to the moon’s standard acceleration gmoon = 1.62 m/s2? (c) How fast will the
body accelerate if a net force of 400 lbf is applied to it on the moon or on the earth?
Solution
We need to find the (a) mass; (b) weight on the moon; and (c) acceleration of this body.
This is a fairly simple example of conversion factors for differing unit systems. No property data is needed. The example is too low level for a sketch.
Part (a)
Newton’s law (1.2) holds with known weight and gravitational acceleration. Solve for m:
F = W = 1000 lbf = mg = (m) (32.174 ft/s2 ),
Convert this to kilograms:
or
m=
1000 lbf
= 31.08 slugs
32.174 ft/s2
m = 31.08 slugs = (31.08 slugs) (14.5939 kg/slug) = 454 kg
Ans. (a)
Part (b)
The mass of the body remains 454 kg regardless of its location. Equation (1.2) applies
with a new gravitational acceleration and hence a new weight:
F = Wmoon = mgmoon = (454 kg) (1.62 m/s2 ) = 735 N = 165 lbf
Ans. (b)
Part (c)
This part does not involve weight or gravity or location. It is simply an application of
­Newton’s law with a known mass and known force:
F = 400 lbf = ma = (31.08 slugs) a
1.4 Dimensions and Units 11
Solve for
a=
400 lbf
ft
m
m
= 12.87 2 ( 0.3048 ) = 3.92 2 31.08 slugs
ft
s
s
Ans. (c)
Comment (c): This acceleration would be the same on the earth or moon or anywhere.
Many data in the literature are reported in inconvenient or arcane units suitable
only to some industry or specialty or country. The engineer should convert these
data to the SI or BG system before using them. This requires the systematic
application of conversion factors, as in the following example.
EXAMPLE 1.2
Industries involved in viscosity measurement [27, 29] continue to use the CGS system
of units, since centimeters and grams yield convenient numbers for many fluids. The
absolute viscosity (µ) unit is the poise, named after J. L. M. Poiseuille, a French physician who in 1840 performed pioneering experiments on water flow in pipes; 1 poise
= 1 g/(cm-s). The kinematic viscosity (ν) unit is the stokes, named after G. G. Stokes,
a British physicist who in 1845 helped develop the basic partial differential equations of
fluid momentum; 1 stokes = 1 cm2/s. Water at 20°C has µ ≈ 0.01 poise and also ν ≈
0.01 stokes. Express these results in (a) SI and (b) BG units.
Solution
Part (a)
∙ Approach: Systematically change grams to kg or slugs and change centimeters to
meters or feet.
∙ Property values: Given µ = 0.01 g/(cm-s) and ν = 0.01 cm2/s.
∙ Solution steps: (a) For conversion to SI units,
g
g(1 kg/1000 g)
kg
μ = 0.01 cm · s = 0.01
= 0.001 m · s
cm(0.01 m/cm)s
ν = 0.01
cm2 (0.01 m/cm) 2
cm2
m2
= 0.01
=
0.000001
s
s
s
Ans. (a)
Part (b)
∙ For conversion to BG units
μ = 0.01
g
g(1 kg/1000 g) (1 slug/14.5939 kg)
slug
= 0.01
= 0.0000209
cm · s
(0.01 m/cm) (1 ft/0.3048 m)s
ft · s
ν = 0.01
cm2 (0.01 m/cm) 2 (1 ft/0.3048 m) 2
cm2
ft2
= 0.01
= 0.0000108 s
s
s
Ans. (b)
∙ Comments: This was a laborious conversion that could have been shortened by
using the direct viscosity conversion factors in App. C or the inside front cover. For
example, µBG = µSI/47.88.
12
Chapter 1 Introduction
We repeat our advice: Faced with data in unusual units, convert them immediately to either SI or BG units because (1) it is more professional and (2) theoretical equations in fluid mechanics are dimensionally consistent and require no
further conversion factors when these two fundamental unit systems are used, as
the following example shows.
EXAMPLE 1.3
A useful theoretical equation for computing the relation between pressure, velocity, and
altitude in a steady flow of a nearly inviscid, nearly incompressible fluid with negligible heat transfer and shaft work5 is the Bernoulli relation, named after Daniel
­Bernoulli, who published a hydrodynamics textbook in 1738:
p0 = p + 12 ρV2 + ρgZ
(1)
where p0 = stagnation pressure
p = pressure in moving fluid
V = velocity
ρ = density
Z = altitude
g = gravitational acceleration
(a) Show that Eq. (1) satisfies the principle of dimensional homogeneity, which states
that all additive terms in a physical equation must have the same dimensions. (b) Show
that consistent units result without additional conversion factors in SI units. (c) Repeat
(b) for BG units.
Solution
Part (a)
We can express Eq. (1) dimensionally, using braces, by entering the dimensions of each
term from Table 1.2:
{ML −1T −2 } = {ML −1T −2 } + {ML −3 }{L2T −2 } + {ML −3 }{LT −2 }{L}
= {ML −1T −2 } for all terms
Ans. (a)
Part (b)
Enter the SI units for each quantity from Table 1.2:
{N/m2 } = {N/m2 } + {kg/m3 }{m2/s2 } + {kg/m3 }{m/s2 }{m}
= {N/m2 } + {kg/(m · s2 ) }
The right-hand side looks bad until we remember from Eq. (1.3) that 1 kg = 1 N · s2/m.
{kg/(m · s2 ) } =
5
{N · s2/m}
= {N/m2 }
{m · s2 }
That’s an awful lot of assumptions, which need further study in Chap. 3.
Ans. (b)
1.4 Dimensions and Units 13
Thus all terms in Bernoulli’s equation will have units of pascals, or newtons per square
meter, when SI units are used. No conversion factors are needed, which is true of all
theoretical equations in fluid mechanics.
Part (c)
Introducing BG units for each term, we have
{lbf/ft2 } = {lbf/ft2 } + {slugs/ft3 }{ft2/s2 } + {slugs/ft3 }{ft/s2 }{ft}
= {lbf/ft2 } + {slugs/(ft · s2 ) }
But, from Eq. (1.3), 1 slug = 1 lbf · s2/ft, so that
{slugs/(ft · s2 ) } =
{lbf · s2/ft}
= {lbf/ft2 }
{ft · s2 }
Ans. (c)
All terms have the unit of pounds-force per square foot. No conversion factors are
needed in the BG system either.
There is still a tendency in English-speaking countries to use pound-force per
square inch as a pressure unit because the numbers are more manageable. For
example, standard atmospheric pressure is 14.7 lbf/in2 = 2116 lbf/ft2 = 101,300
Pa. The pascal is a small unit because the newton is less than 14 lbf and a square
meter is a very large area.
Consistent Units
Note that not only must all (fluid) mechanics equations be dimensionally homogeneous, one must also use consistent units; that is, each additive term must have
the same units. There is no trouble doing this with the SI and BG systems, as
in Example 1.3, but woe unto those who try to mix colloquial English units. For
example, in Chap. 9, we often use the assumption of steady adiabatic compressible gas flow:
h + 12V 2 = constant
where h is the fluid enthalpy and V2/2 is its kinetic energy per unit mass. Colloquial thermodynamic tables might list h in units of British thermal units per
pound mass (Btu/lb), whereas V is likely used in ft/s. It is completely erroneous
to add Btu/lb to ft2/s2. The proper unit for h in this case is ft · lbf/slug, which is
identical to ft2/s2. The conversion factor is 1 Btu/lb ≈ 25,040 ft2/s2 = 25,040
ft · lbf/slug.
Homogeneous versus Dimensionally Inconsistent Equations
All theoretical equations in mechanics (and in other physical sciences) are
­dimensionally homogeneous; that is, each additive term in the equation has the
14
Chapter 1 Introduction
same dimensions. However, the reader should be warned that many empirical
formulas in the ­engineering literature, arising primarily from correlations of data,
are dimensionally inconsistent. Their units cannot be reconciled simply, and some
terms may contain hidden variables. An example is the formula that pipe valve
manufacturers cite for liquid volume flow rate Q (m3/s) through a partially open
valve:
Q = CV (
Δp 1/2
SG )
where Δp is the pressure drop across the valve and SG is the specific gravity of the liquid (the ratio of its density to that of water). The quantity CV
is the valve flow coefficient, which manufacturers tabulate in their valve
brochures. Since SG is dimensionless {1}, we see that this formula is totally
inconsistent, with one side being a flow rate {L3/T} and the other being the
square root of a pressure drop {M1/2/L1/2T}. It follows that CV must have
dimensions, and rather odd ones at that: {L7/2/M1/2}. Nor is the resolution of
this discrepancy clear, although one hint is that the values of CV in the literature increase nearly as the square of the size of the valve. The presentation
of experimental data in homogeneous form is the subject of dimensional
analysis (Chap. 5). There we shall learn that a homogeneous form for the
valve flow relation is
Q = Cd Aopening(
Table 1.3 Convenient Prefixes
for Engineering Units
Multiplicative
factor
12
10
109
106
103
102
10
10−1
10−2
10−3
10−6
10−9
10−12
10−15
10−18
Prefix
Symbol
tera
giga
mega
kilo
hecto
deka
deci
centi
milli
micro
nano
pico
femto
atto
T
G
M
k
h
da
d
c
m
µ
n
p
f
a
Δp 1/2
ρ )
where ρ is the liquid density and A the area of the valve opening. The discharge
coefficient Cd is dimensionless and changes only moderately with valve size.
Please believe—until we establish the fact in Chap. 5—that this latter is a much
better formulation of the data.
Meanwhile, we conclude that dimensionally inconsistent equations, though
they occur in engineering practice, are misleading and vague and even
­dangerous, in the sense that they are often misused outside their range of
applicability.
Convenient Prefixes in Powers of 10
Engineering results often are too small or too large for the common units, with
too many zeros one way or the other. For example, to write p = 114,000,000 Pa
is long and awkward. Using the prefix “M” to mean 106, we convert this to a
concise p = 114 MPa (megapascals). Similarly, t = 0.000000003 s is a proofreader’s nightmare compared to the equivalent t = 3 ns (nanoseconds). Such
prefixes are common and convenient, in both the SI and BG systems. A complete
list is given in Table 1.3.
1.4 Dimensions and Units 15
EXAMPLE 1.4
In 1890 Robert Manning, an Irish engineer, proposed the following empirical formula
for the average velocity V in uniform flow due to gravity down an open channel (BG
units):
V=
1.49 2/3 1/2
R S (1)
n
where R = hydraulic radius of channel (Chaps. 6 and 10)
S = channel slope (tangent of angle that bottom makes with horizontal)
n = Manning’s roughness factor (Chap. 10)
and n is a constant for a given surface condition for the walls and bottom of the channel. (a) Is Manning’s formula dimensionally consistent? (b) Equation (1) is commonly
taken to be valid in BG units with n taken as dimensionless. Rewrite it in SI form.
Solution
∙ Assumption: The channel slope S is the tangent of an angle and is thus a dimensionless ratio with the dimensional notation {1}—that is, not containing M, L, or T.
∙ Approach (a): Rewrite the dimensions of each term in Manning’s equation, using
­brackets {}:
1.49
{V} = {
{R2/3 } {S1/2 }
n }
or
L
1.49
2/3
{T} = { n }{L } {1}
This formula is incompatible unless {1.49/n} = {L1/3/T}. If n is dimensionless (and
it is never listed with units in textbooks), the number 1.49 must carry the dimensions
of {L1/3/T}.
Ans. (a)
∙ Comment (a): Formulas whose numerical coefficients have units can be disastrous
for engineers working in a different system or another fluid. Manning’s formula,
though popular, is inconsistent both dimensionally and physically and is valid only
for water flow with certain wall roughnesses. The effects of water viscosity and
density are hidden in the numerical value 1.49.
∙ Approach (b): Part (a) showed that 1.49 has dimensions. If the formula is valid in
BG units, then it must equal 1.49 ft1/3/s. By using the SI conversion for length, we
obtain
(1.49 ft1/3/s) (0.3048 m/ft) 1/3 = 1.00 m1/3/s
Therefore, Manning’s inconsistent formula changes form when converted to the SI
system:
SI units:
V=
1.0 2/3 1/2
R S n
Ans. (b)
with R in meters and V in meters per second.
∙ Comment (b): Actually, we misled you: This is the way Manning, a metric user,
first proposed the formula. It was later converted to BG units. Such dimensionally
inconsistent formulas are dangerous and should be either reanalyzed or treated as
having very limited application.
16
Chapter 1 Introduction
1.5 System and Control Volume
A system is defined as an arbitrary quantity of mass of fixed identity. It is also
called closed system or control mass. Everything external to the system is denoted
by the term surroundings. The system is separated from the surroundings by its
boundaries. The boundaries of the system can be fixed or movable. Fig. 1.3a
shows a piston-cylinder assembly. The gas in the cylinder can be considered the
system. If the piston is pushed by an external force, the gas is compressed; the
boundary of the system thus moves. Heat and work may cross the boundaries of
the system, but the quantity of the mass in the system remains unchanged. The
concept of the system is almost exclusively used in solid mechanics, where we
are more likely to be interested in the trajectory of a particle or a rigid body for
analysis.
On the other hand, in thermodynamics and fluid mechanics, when fluid flow
is of the interest of the study, one can rarely keep track of the actual fate of the
specific fluid particles, as they are numerous in quantity and easily deformable. A
control volume is defined as an arbitrary region in space through which fluid flows.
It can be at rest or in motion. The geometric boundaries of the control volume
separating it from the surrounding is called control surface. The control surface
can be real or imaginary. Fig. 1.3b illustrates flow through a pipe T-junction. A
control surface is drawn on the junction. The area of the control surface corresponding to the walls of pipe are the physical and real surfaces, whereas others
representing inlets or outlets are the imaginary surfaces.
Mathematically, a boundary or a control surface has zero thickness, and thus
it can neither contain any mass nor occupy any volume in space.
Methods of Description
The type of analysis for a physical system depends on the problem. In solid
mechanics it is easy to keep track of a fixed and identifiable quantity of mass,
such as particles or rigid bodies. A frame of reference that follows the trajectories
of individual particles or bodies is used and referred to as the Lagrangian method
of description. Most of the basic physical laws are written for a system of mass
(Chap. 3). The analysis based on the Lagrangian method requires us to track the
position, velocity, and acceleration of each individual particle.
V1
Moving boundary
Imaginary surface
Fig. 1.3 System and control volume:
(a) gas in a closed system with a
moving boundary; (b) fluid flow
through a control volume with real
and imaginary surfaces.
F
Gas
V3
Control volume
Real surface
Fixed boundary
(a)
(b)
V2
1.5 System and Control Volume 17
In a situation of fluid flow, however, the determination, by experiment or
theory, of the properties of the fluid as a function of position and time is
considered to be the solution to the problem. This is because it would not
be possible to keep track of the motion of each particle of fluid. Consequently, a Lagrangian description of particles would be unmanageable. Often
we find it convenient to adopt a different method of description—the Eulerian method, as we commonly call it. Particularly with analyses of control
volume, the emphasis is on the space-time distribution of the fluid properties.
In the Eulerian method of description, the properties of a flow field are
described as functions of space coordinates and time. This treatment of properties as a logical outgrowth of the continuum-field assumption distinguishes
fluid mechanics from solid mechanics. The properties of the velocity and
acceleration fields are demonstrated below using the Eulerian method of
description.
The Velocity Field
Foremost among the properties of a flow is the velocity field V(x, y, z, t). In fact,
determining the velocity is often tantamount to solving a flow problem, since
other properties follow directly from the velocity field. Chapter 2 is devoted to
the calculation of the pressure field once the velocity field is known. Books on
heat transfer (for example, Ref. 20) are largely devoted to finding the temperature
field from known velocity fields.
In general, velocity is a vector function of position and time and thus has three
components u, v, and w, each a scalar field in itself:
V( x, y, z, t) = iu(x, y, z, t) + jv(x, y, z, t) + kw(x, y, z, t) (1.4)
The use of u, v, and w instead of the more logical component notation Vx, Vy,
and Vz is the result of an almost unbreakable custom in fluid mechanics. Much
of this textbook, especially Chaps. 4, 7, 8, and 9, is concerned with finding the
distribution of the velocity vector V for a variety of practical flows.
The Acceleration Field
The acceleration vector, a = dV/dt, occurs in Newton’s law for a fluid and thus
is very important. In order to follow a particle in the Eulerian frame of reference,
the final result for acceleration is nonlinear and quite complicated. Here we only
give the formula:
a=
dV ∂V
∂V
∂V
∂V
=
+u
+v
+w
dt
∂t
∂x
∂y
∂z
(1.5)
where (u, v, w) are the velocity components from Eq. (1.4). We shall study this
formula in detail in Chap. 4. The last three terms in Eq. (1.5) are nonlinear
­products and greatly complicate the analysis of general fluid motions, especially
viscous flows.
18
Chapter 1 Introduction
1.6 Thermodynamic Properties of a Fluid
While the velocity field V is the most important fluid property, it interacts closely
with the thermodynamic properties of the fluid. We have already introduced into
the discussion the three most common such properties:
1. Pressure p
2. Density ρ
3. Temperature T
These three are constant companions of the velocity vector in flow analyses. Four
other intensive thermodynamic properties become important when work, heat,
and energy balances are treated (Chaps. 3 and 4):
4.
5.
6.
7.
Internal energy û
Enthalpy h = û + p/ρ
Entropy s
Specific heats cp and cv
In addition, friction and heat conduction effects are governed by the two so-called
transport properties:
8. Coefficient of viscosity µ
9. Thermal conductivity k
All nine of these quantities are true thermodynamic properties that are determined
by the thermodynamic condition or state of the fluid. For example, for a singlephase substance such as water or oxygen, two basic properties such as pressure
and temperature are sufficient to fix the value of all the others:
ρ = ρ(p, T)
h = h(p, T)
μ = μ(p, T)
and so on for every quantity in the list. Note that the specific volume, so important in thermodynamic analyses, is omitted here in favor of its inverse, the
­density ρ.
Recall that thermodynamic properties describe the state of a system—that is,
a collection of matter of fixed identity that interacts with its surroundings. In
most cases here the system will be a small fluid element, and all properties will
be assumed to be continuum properties of the flow field: ρ = ρ(x, y, z, t), and
so on.
Recall also that thermodynamics is normally concerned with static systems,
whereas fluids are usually in variable motion with constantly changing properties.
Do the properties retain their meaning in a fluid flow that is technically not in
equilibrium? The answer is yes, from a statistical argument. In gases at normal
pressure (and even more so for liquids), an enormous number of molecular collisions occur over a very short distance of the order of 1 µm, so that a fluid
subjected to sudden changes rapidly adjusts itself toward equilibrium. We therefore assume that all the thermodynamic properties just listed exist as point functions in a flowing fluid and follow all the laws and state relations of ordinary
equilibrium thermodynamics. There are, of course, important nonequilibrium
1.6 Thermodynamic Properties of a Fluid 19
effects such as chemical and nuclear reactions in flowing fluids, which are not
treated in this text.
Pressure
Pressure is the (compression) stress at a point in a static fluid (Fig. 1.1). Next to
velocity, the pressure p is the most dynamic variable in fluid mechanics. Differences or gradients in pressure often drive a fluid flow, especially in ducts. In
low-speed flows, the actual magnitude of the pressure is often not important,
unless it drops so low as to cause vapor bubbles to form in a liquid. For convenience, we set many such problem assignments at the level of 1 atm = 2116 lbf/
ft2 = 101,300 Pa. High-speed (compressible) gas flows (Chap. 9), however, are
indeed sensitive to the magnitude of pressure.
Temperature
Temperature T is related to the internal energy level of a fluid. It may vary considerably during high-speed flow of a gas (Chap. 9). Although engineers often
use Celsius or Fahrenheit scales for convenience, many applications in this text
require absolute (Kelvin or Rankine) temperature scales:
°R = °F + 459.69
K = °C + 273.16
If temperature differences are strong, heat transfer may be important [20], but
our concern here is mainly with dynamic effects.
Density
The density of a fluid, denoted by ρ (lowercase Greek rho), is its mass per unit
­volume. Density is highly variable in gases and increases nearly proportionally to
the pressure level. Density in liquids is nearly constant; the density of water (about
1000 kg/m3) increases only 1 percent if the pressure is increased by a factor of
220. Thus most liquid flows are treated analytically as nearly “incompressible.”
In general, liquids are about three orders of magnitude more dense than gases
at atmospheric pressure. The heaviest common liquid is mercury, and the lightest
gas is hydrogen. Compare their densities at 20°C and 1 atm:
Mercury: ρ = 13,580 kg/m3
Hydrogen: ρ = 0.0838 kg/m3
They differ by a factor of 162,000! Thus, the physical parameters in various
liquid and gas flows might vary considerably. The differences are often resolved
by the use of dimensional analysis (Chap. 5). Other fluid densities are listed in
Tables A.3 and A.4 (in App. A), and in Ref. 21.
Specific Weight
The specific weight of a fluid, denoted by γ (lowercase Greek gamma), is its
weight per unit volume. Just as a mass has a weight W = mg, density and specific
weight are simply related by gravity:
γ = ρg
(1.6)
20
Chapter 1 Introduction
The units of γ are weight per unit volume, in lbf/ft3 or N/m3. In standard earth
gravity, g = 32.174 ft/s2 = 9.807 m/s2. Thus, for example, the specific weights
of air and water at 20°C and 1 atm are approximately
γair = (1.205 kg/m3 )(9.807 m/s2 ) = 11.8 N/m3 = 0.0752 lbf/ft3
γwater = (998 kg/m3 )(9.807 m/s2 ) = 9790 N/m3 = 62.4 lbf/ft3
Specific weight is very useful in the hydrostatic pressure applications of Chap. 2.
Specific weights of other fluids are given in Tables A.3 and A.4.
Specific Gravity
Specific gravity, denoted by SG, is the ratio of a fluid density to a standard reference fluid, usually water at 4°C (for liquids) and air (for gases):
SGgas =
SGliquid =
ρgas
ρgas
=
ρair 1.205 kg/m3
ρliquid
ρwater
=
(1.7)
ρliquid
1000 kg/m3
For example, the specific gravity of mercury (Hg) is SGHg = 13,580/1000 ≈ 13.6.
Engineers find these dimensionless ratios easier to remember than the actual
numerical values of density of a variety of fluids.
Potential and Kinetic Energies
In thermostatics the only energy in a substance is that stored in a system by
molecular activity and molecular bonding forces. This is commonly denoted as
internal energy û. A commonly accepted adjustment to this static situation for
fluid flow is to add two more energy terms that arise from newtonian mechanics:
potential energy and kinetic energy.
The potential energy equals the work required to move the system of mass m
from the origin to a position vector r = ix + jy + kz against a gravity field g.
Its value is −mg · r, or −g · r per unit mass. The kinetic energy equals the work
required to change the speed of the mass from zero to velocity V. Its value is
1
1 2
2
2 mV or 2 V per unit mass. Then by common convention the total stored energy
e per unit mass in fluid mechanics is the sum of three terms:
e = û + 12V 2 + (–g · r) (1.8)
Also, throughout this book we shall define z as upward, so that g = −gk and
g · r = −gz. Then Eq. (1.8) becomes
e = û + 12V 2 + gz
(1.9)
The molecular internal energy û is a function of T and p for the single-phase
pure substance, whereas the potential and kinetic energies are kinematic
­quantities.
1.6 Thermodynamic Properties of a Fluid 21
State Relations for Gases
Thermodynamic properties are found both theoretically and experimentally to be
related to each other by state relations that differ for each substance. As mentioned, we shall confine ourselves here to single-phase pure substances, such as
water in its liquid phase. The second most common fluid, air, is a mixture of
gases, but since the mixture ratios remain nearly constant between 160 and 2200
K, in this temperature range air can be considered to be a pure substance.
All gases at high temperatures and low pressures (relative to their critical
point) are in good agreement with the perfect-gas law
p = ρRT R = cp − cv = gas constant (1.10)
where the specific heats cp and cv are defined in Eqs. (1.14) and (1.15).
Since Eq. (1.10) is dimensionally consistent, R has the same dimensions as
specific heat, {L2T−2Θ−1}, or velocity squared per temperature unit (kelvin or
degree Rankine). Each gas has its own constant R, equal to a universal constant
Λ divided by the molecular weight
Rgas =
Λ
Mgas
(1.11)
where Λ = 49,700 ft-lbf/(slugmol · °R) = 8314 J/(kmol · K). Most applications in
this book are for air, whose molecular weight is M = 28.97g/mol = 28.97 lb/lbmol:
49,700 ft · lbf/(slugmol · °R)
m2
ft · lbf
ft2
= 287 2
= 1716
= 1716 2
28.97lb/lbmol
slug · °R
s ·K
s °R
(1.12)
Rair =
Standard atmospheric pressure is 2116 lbf/ft2 = 2116 slug/(ft · s2), and standard
temperature is 60°F = 520°R. Thus standard air density is
ρair =
2116 slug/(ft · s2 )
= 0.00237 slug/ft3 = 1.22 kg/m3
[1716 ft2/(s2 · °R)](520°R)
(1.13)
This is a nominal value suitable for problems. For other gases, see Table A.4.
Most of the common gases—oxygen, nitrogen, hydrogen, helium, argon—are
nearly ideal. This is not so true for steam, whose simplified temperature–entropy
chart is shown in Fig. 1.4. Unless you are sure that the steam temperature is
“high” and the pressure “low,” it is best to use the steam tables to make accurate
calculations.
One proves in thermodynamics that Eq. (1.10) requires that the internal molecular energy û of a perfect gas vary only with temperature: û = û(T ). Therefore,
the specific heat cv also varies only with temperature:
∂u^ �
du^ �
cv = (
=
= cv (T )
)
∂T ρ
dT
or
du^ �= cv (T )dT
(1.14)
22
Chapter 1 Introduction
800
700
p = 22,060 kPa
= 10,000
= 1,000
= 100
= 10
Fig. 1.4 Temperature–entropy chart
for steam. The critical point is pc =
22,060 kPa, Tc = 374°C, Sc =
4.41 kJ/(kg · K). Except near the
critical point, the smooth isobars
tempt one to assume, often incorrectly, that the ideal-gas law is valid
for steam. It is not, except at low
pressure and high temperature: the
upper right of the graph.
Temperature, °C
600
500
400
Critical point
300
200
Two-phase region
100
0
1
2
3
4
5
6
7
.
Entropy, kJ/(kg K)
8
9
10
11
In like manner the enthalpy h and the specific heat cp of a perfect gas also vary
only with temperature:
p
h=�
u^ + = �
u^ + RT = h(T )
ρ
∂h
dh
cp = ( ) =
= cp (T )
∂T p dT
dh = cp ( T )dT
(1.15)
The ratio of specific heats of a perfect gas is an important dimensionless parameter in compressible flow analysis (Chap. 9)
k=
cp
= k(T ) ≥ 1
cv
(1.16)
As a first approximation in airflow analysis we commonly take cp, cv, and k to
be constant:
kair ≈ 1.4
R
cv =
≈ 4293 ft2/(s2 · °R) = 718 m2/(s2 · K)
k−1
kR
cp =
≈ 6010 ft2/(s2 · °R) = 1005 m2/(s2 · K)
k−1
(1.17)
Actually, for all gases, cp and cv increase gradually with temperature, and k
decreases gradually. Experimental values of the specific-heat ratio for eight common gases are shown in Fig. 1.5. Nominal values are in Table A.4.
Many flow problems involve steam. Typical steam operating conditions are
often close to the critical point, so that the perfect-gas approximation is inaccurate. Then we must turn to the steam tables, either in tabular or CD-ROM form
[23] or as online software [24]. Most online steam tables require a license fee,
1.6 Thermodynamic Properties of a Fluid 23
1.7
Ar
1.6
Atmospheric pressure
1.5
H2
1.4
cp
k= c
υ
CO
1.3
O2
Air and
N2
Steam
1.2
CO2
1.1
Fig. 1.5 Specific-heat ratio of eight
common gases as a function of
­temperature. (Data from Ref. 22.)
1.0
0
1000
2000
3000
4000
5000
Temperature, °R
but the writer, in ­Example 1.5 that follows, suggests a free online source. Sometimes the error of using the perfect-gas law for steam can moderate, as the
­following ­example shows.
EXAMPLE 1.5
Estimate ρ and cp of steam at 100 lbf/in2 and 400°F, in English units, (a) by the perfectgas approximation and (b) by the ASME Steam Tables [23].
Solution
∙ Approach (a)—the perfect-gas law: Although steam is not an ideal gas, we can
estimate these properties with moderate accuracy from Eqs. (1.10) and (1.17). First
convert pressure from 100 lbf/in2 to 14,400 lbf/ft2, and use absolute temperature,
(400°F + 460) = 860°R. Then we need the gas constant for steam, in English units.
From Table A.4, the ­molecular weight of H2O is 18.02, whence
Rsteam =
ΛEnglish
MH2O
=
49,700 ft · lbf/(slugmol °R)
ft · lbf
= 2758
18.02/mol
slug °R
24
Chapter 1 Introduction
Then the density estimate follows from the perfect-gas law, Eq. (1.10):
ρ≈
slug
p
14,400 lbf/ft2
=
≈ 0.00607 3 RT [2758 ft · lbf/(slug · °R) ] (860 °R)
ft
Ans. (a)
At 860°R, from Fig. 1.5, ksteam = cp/cv ≈ 1.30. Then, from Eq. (1.17),
cp ≈
(1.3) (2758 ft · lbf/(slug °R) )
kR
ft · lbf
=
≈ 12,000
k−1
(1.3 − 1)
slug °R
Ans. (a)
∙ Approach (b)—tables or software: One can either read the ASME Steam Tables
[23] or use online software. Online software, such as [24], calculates the properties
of steam without reading a table. Most of these require a license fee, which your
institution may or may not possess. For work at home, the writer has found success
with this free commercial online site:
https://beta.spiraxsarco.com/resources-and-design-tools
The software calculates superheated steam properties, as required in this example. The
Spirax Sarco Company makes many types of steam equipment: boilers, condensers,
valves, pumps, regulators. This site provides many steam properties—density, specific
heat, enthalpy, speed of sound—in many different unit systems. Here we need the
density and specific heat of steam at 100 lbf/in2 and 400°F. You enter these two inputs
and it will calculate not only ρ and cp but also many other properties of interest, in
English or metric units. The software results are:
ρ(100 lbf/in2, 400°F) = 0.2027 lbm/ft3 = 3.247 kg/m3
Ans. (b)
cp (100 lbf/in , 400°F) = 0.5289 Btu/(lbm-F) = 2215 J/(kg-K)
Ans. (b)
2
Comments: These are quite accurate and compare well to other steam tables. The perfectgas estimate of ρ is 4 percent low, and the estimate of cp is 9 percent low. The chief reason
for the discrepancy is that this temperature and pressure are rather close to the critical point
and saturation line of steam. At higher temperatures and lower pressures, say, 800°F and
50 lbf/in2, the perfect-gas law has an accuracy of about ±1 percent. See Fig. 1.4.
Again let us warn that English units (psia, lbm, Btu) are awkward and must be
converted to SI or BG units in almost all fluid mechanics formulas.
State Relations for Liquids
The writer knows of no “perfect-liquid law” comparable to that for gases. Liquids
are nearly incompressible and have a single, reasonably constant specific heat.
Thus an idealized state relation for a liquid is
ρ ≈ const cp ≈ cv ≈ const dh ≈ cp dT
(1.18)
Most of the flow problems in this book can be attacked with these simple assumptions. Water is normally taken to have a density of 998 kg/m3 and a specific heat
cp = 4210 m2/(s2 · K). The steam tables may be used if more accuracy is required.
The density of a liquid usually decreases slightly with temperature and increases
moderately with pressure. If we neglect the temperature effect, an empirical pressure–density relation for a liquid is
p
ρ n
≈ (B + 1) ( ) − B
pa
ρa
(1.19)
1.7 Viscosity and Other Secondary Properties 25
where B and n are dimensionless parameters that vary slightly with temperature
and pa and ρa are standard atmospheric values. Water can be fitted approximately
to the values B ≈ 3000 and n ≈ 7.
Seawater is a variable mixture of water and salt and thus requires three thermodynamic properties to define its state. These are normally taken as pressure,
temperature, and the salinity Ŝ, defined as the weight of the dissolved salt divided
by the weight of the mixture. The average salinity of seawater is 0.035, usually
written as 35 parts per 1000, or 35 ‰. The average density of seawater is 2.00
slugs/ft3 ≈ 1030 kg/m3. Strictly speaking, seawater has three specific heats, all
approximately equal to the value for pure water of 25,200 ft2/(s2 · °R) = 4210
m2/(s2 · K).
EXAMPLE 1.6
The pressure at the deepest part of the ocean is approximately 1100 atm. Estimate the
density of seawater in slug/ft3 at this pressure.
Solution
Equation (1.19) holds for either water or seawater. The ratio p/pa is given as 1100:
or
ρ 7
1100 ≈ (3001) ( ) − 3000
ρa
ρ
4100 1/7
=(
= 1.046
ρa
3001 )
Assuming an average surface seawater density ρa = 2.00 slugs/ft3, we compute
ρ ≈ 1.046(2.00) = 2.09 slugs/ft3
Ans.
Even at these immense pressures, the density increase is less than 5 percent, which
justifies the treatment of a liquid flow as essentially incompressible.
1.7 Viscosity and Other Secondary Properties
The quantities such as pressure, temperature, and density discussed in the previous section are primary thermodynamic variables characteristic of any system.
Certain secondary variables also characterize specific fluid mechanical behavior.
The most important of these is viscosity, which relates the local stresses in a
moving fluid to the strain rate of the fluid element.
Viscosity
Viscosity is a quantitative measure of a fluid’s resistance to flow. More specifically, it determines the fluid strain rate that is generated by a given applied shear
stress. We can easily move through air, which has very low viscosity. Movement
is more difficult in water, which has 50 times higher viscosity. Still more resistance is found in SAE 30 oil, which is 300 times more viscous than water. Try
26
Chapter 1 Introduction
to slide your hand through glycerin, which is five times more viscous than SAE
30 oil, or blackstrap molasses, another factor of five higher than glycerin. Fluids
may have a vast range of viscosities.
Consider a fluid element sheared in one plane by a single shear stress τ, as in
Fig. 1.6a. The shear strain angle δθ will continuously grow with time as long as
the stress τ is maintained, the upper surface moving at speed δu larger than the
lower. Such common fluids as water, oil, and air show a linear relation between
applied shear and resulting strain rate:
τ ∝
δθ
δt
(1.20)
From the geometry of Fig. 1.6a, we see that
tan δθ =
δu δt
δy
(1.21)
In the limit of infinitesimal changes, this becomes a relation between shear strain
rate and velocity gradient:
dθ du
= dt
dy
(1.22)
From Eq. (1.20), then, the applied shear is also proportional to the velocity
gradient for the common linear fluids. The constant of proportionality is the
viscosity coefficient µ:
τ=μ
dθ
du
=μ
dt
dy
(1.23)
Equation (1.23) is dimensionally consistent; therefore µ has dimensions of stress–
time: {FT/L2} or {M/(LT )}. The BG unit is slugs per foot-second, and the SI unit
is kilograms per meter-second. The linear fluids that follow Eq. (1.23) are called
y
δu δt
τ∝
u( y)
δθ
δt
u = δu
δθ
du
δθ
τ = µ du
dy
dy
δy
Fig. 1.6 Shear stress causes continuous shear deformation in a fluid:
(a) a fluid element straining at a rate
δθ/δt; (b) newtonian shear distribution in a shear layer near a wall.
δx
u=0
τ
(a)
0
Velocity
profile
No slip at wall
(b)
1.7 Viscosity and Other Secondary Properties 27
Table 1.4 Viscosity and Kinematic Viscosity of Eight Fluids at 1 atm and 20°C
Fluid
Hydrogen
Air
Gasoline
Water
Ethyl alcohol
Mercury
SAE 30 oil
Glycerin
µ,
kg/(m · s)†
Ratio
µ/µ(H2)
ρ,
kg/m3
9.0 E–6
1.8 E–5
2.9 E–4
1.0 E–3
1.2 E–3
1.5 E–3
0.29
1.5
1.0
2.1
33
114
135
170
33,000
170,000
0.084
1.20
680
998
789
13,550
891
1,260
ν
m2/s†
1.05
1.50
4.22
1.01
1.52
1.16
3.25
1.18
E–4
E–5
E–7
E–6
E–6
E–7
E–4
E–3
Ratio
ν/ν(Hg)
910
130
3.7
8.7
13
1.0
2,850
10,300
†
1 kg/(m · s) = 0.0209 slug/(ft · s); 1 m2/s = 10.76 ft2/s.
newtonian fluids, after Sir Isaac Newton, who first postulated this resistance law
in 1687.
We do not really care about the strain angle θ(t) in fluid mechanics, concentrating instead on the velocity distribution u(y), as in Fig. 1.6b. We shall
use Eq. (1.23) in Chap. 4 to derive a differential equation for finding the
velocity distribution u(y)—and, more generally, V(x, y, z, t)—in a viscous
fluid. Figure 1.6b illustrates a shear layer, or boundary layer, near a solid
wall. The shear stress is proportional to the slope of the velocity profile and
is greatest at the wall. Further, at the wall, the velocity u is zero relative to
the wall: This is called the no-slip condition and is characteristic of all viscous
fluid flows.
The viscosity of newtonian fluids is a true thermodynamic property and varies
with temperature and pressure. At a given state (p, T ) there is a vast range of
values among the common fluids. Table 1.4 lists the viscosity of eight fluids at
standard pressure and temperature. There is a variation of six orders of magnitude
from hydrogen up to glycerin. Thus there will be wide differences between fluids
subjected to the same applied stresses.
Generally speaking, the viscosity of a fluid increases only weakly with pressure. For example, increasing p from 1 to 50 atm will increase µ of air only 10
percent. Temperature, however, has a strong effect, with µ increasing with T for
gases and decreasing for liquids. Figure A.1 (in App. A) shows this temperature
variation for various common fluids. It is customary in most engineering work
to neglect the pressure variation.
The variation µ(p, T ) for a typical fluid is nicely shown by Fig. 1.7, from Ref.
25, which normalizes the data with the critical-point state (µc, pc, Tc). This behavior, called the principle of corresponding states, is characteristic of all fluids, but
the actual numerical values are uncertain to ±20 percent for any given fluid. For
example, values of µ(T ) for air at 1 atm, from Table A.2, fall about 8 percent
low compared to the “low-density limit” in Fig. 1.7.
Note in Fig. 1.7 that changes with temperature occur very rapidly near the
critical point. In general, critical-point measurements are extremely difficult and
uncertain.
28
Chapter 1 Introduction
10
9
8
7
6
Liquid
5
4
Dense gas
3
µ
µr = µ
Two-phase
region
2
25
10
c
1
0.9
0.8
0.7
0.6
0.5
0.4
Fig. 1.7 Fluid viscosity
­nondimensionalized by ­critical-point
properties. This generalized chart is
characteristic of all fluids but is
­accurate only to ±20 percent.
(From Ref. 25.)
5
Critical
point
3
2
1
0.5
pr = p/pc = 0.2
0.3
0.2
0.4
Low-density limit
0
0.6
0.8
1
2
3
4
5
6 7 8 9 10
Tr = T
Tc
The Reynolds Number
The primary parameter correlating the viscous behavior of all newtonian fluids
is the dimensionless Reynolds number:
ρVL VL
=
Re =
(1.24)
μ
ν
where V and L are characteristic velocity and length scales of the flow. The
second form of Re illustrates that the ratio of µ to ρ has its own name, the kinematic ­viscosity:
μ
ν= ρ
(1.25)
It is called kinematic because the mass units cancel, leaving only the dimensions
{L2/T}.
Generally, the first thing a fluids engineer should do is estimate the Reynolds
number range of the flow under study. Very low Re indicates viscous creeping
motion, where inertia effects are negligible. Moderate Re implies a smoothly
varying laminar flow. High Re probably spells turbulent flow, which is slowly
varying in the time-mean but has superimposed strong random high-frequency
1.7 Viscosity and Other Secondary Properties 29
fluctuations. Explicit numerical values for low, moderate, and high Reynolds
numbers cannot be stated here. They depend on flow geometry and will be discussed in Chaps. 5 through 7.
Table 1.4 also lists values of ν for the same eight fluids. The pecking order
changes considerably, and mercury, the heaviest, has the smallest viscosity relative to its own weight. All gases have high ν relative to thin liquids such as
gasoline, water, and alcohol. Oil and glycerin still have the highest ν, but the ratio
is smaller. For given values of V and L in a flow, these fluids exhibit a spread
of four orders of magnitude in the Reynolds number.
Flow between Plates
A classic problem is the flow induced between a fixed lower plate and an upper
plate moving steadily at velocity V, as shown in Fig. 1.8. The clearance between
plates is h, and the fluid is newtonian and does not slip at either plate. If the
plates are large, this steady shearing motion will set up a velocity distribution
u(y), as shown, with v = w = 0. The fluid acceleration is zero everywhere.
With zero acceleration and assuming no pressure variation in the flow direction,
you should show that a force balance on a small fluid element leads to the result
that the shear stress is constant throughout the fluid. Then Eq. (1.23) becomes
du τ
= = const
dy μ
which we can integrate to obtain
u = a + by
The velocity distribution is linear, as shown in Fig. 1.8, and the constants a and
b can be evaluated from the no-slip condition at the upper and lower walls:
0 = a + b (0)
u={
V = a + b (h)
at y = 0
at y = h
Hence a = 0 and b = V/h. Then the velocity profile between the plates is given by
y
u=V h
(1.26)
as indicated in Fig. 1.8. Turbulent flow (Chap. 6) does not have this shape.
y
u=V
V
h
Fig. 1.8 Viscous flow induced by
relative motion between two
parallel plates.
u(y)
u=0
Moving
plate:
u=V
Viscous
fluid
x
Fixed plate
30
Chapter 1 Introduction
Although viscosity has a profound effect on fluid motion, the actual viscous
stresses are quite small in magnitude even for oils, as shown in the next example.
Problem-Solving Procedure
Fluid flow analysis is packed with problems to be solved. The present text has
more than 1700 problem assignments. Solving a large number of these is a key
to learning the subject. One must deal with equations, data, tables, assumptions,
unit systems, and solution schemes. The degree of difficulty will vary, and we
urge you to sample the whole spectrum of assignments, with or without the
answers in the Appendix. Here are the recommended steps for problem solution:
1. Read the problem and restate it with your summary of the results desired.
2. From tables or charts, gather the needed property data: density, viscosity,
etc.
3. Make sure you understand what is asked. Students are apt to answer the
wrong question—for example, pressure instead of pressure gradient, lift
force instead of drag force, or mass flow instead of volume flow. Read the
problem carefully.
4. Make a detailed, labeled sketch of the system or control volume needed.
5. Think carefully and list your assumptions. You must decide if the flow is
steady or unsteady, compressible or incompressible, viscous or inviscid,
and whether a control volume or partial differential equations are needed.
6. Find an algebraic solution if possible. Then, if a numerical value is needed,
use either the SI or BG unit systems reviewed in Sec. 1.4.
7. Report your solution, labeled, with the proper units and the proper number
of significant figures (usually two or three) that the data uncertainty allows.
We shall follow these steps, where appropriate, in our example problems.
EXAMPLE 1.7
Suppose that the fluid being sheared in Fig. 1.8 is SAE 30 oil at 20°C. Compute the
shear stress in the oil if V = 3 m/s and h = 2 cm.
Solution
∙ System sketch: This is shown earlier in Fig. 1.8.
∙ Assumptions: Linear velocity profile, laminar newtonian fluid, no slip at either plate
surface.
∙ Approach: The analysis of Fig. 1.8 leads to Eq. (1.26) for laminar flow.
∙ Property values: From Table 1.4 for SAE 30 oil, the oil viscosity µ = 0.29 kg/(m-s).
∙ Solution steps: In Eq. (1.26), the only unknown is the fluid shear stress:
τ=μ
kg
kg · m/s2
(3 m/s)
V
N
= (0.29
=
43.5
= 43.5 2 ≈ 44Pa
2
m · s ) (0.02 m)
h
m
m
Ans.
1.7 Viscosity and Other Secondary Properties 31
∙ Comments: Note the unit identities, 1 kg-m/s2 ≡ 1 N and 1 N/m2 ≡ 1 Pa. Although
oil is very viscous, this shear stress is modest, about 2400 times less than atmospheric
pressure. Viscous stresses in gases and thin (watery) liquids are even smaller.
Variation of Viscosity with Temperature
Temperature has a strong effect and pressure a moderate effect on viscosity. The
viscosity of gases and most liquids increases slowly with pressure. Water is anomalous in showing a very slight decrease below 30°C. Since the change in viscosity
is only a few percent up to 100 atm, we shall neglect pressure effects in this book.
Gas viscosity increases with temperature. Two common approximations are
the power law and the Sutherland law:
T n
( T0 )
μ
≈ µ
μ0
(T/T0 ) 3/2 (T0 + S)
T+S
power law
(1.27)
Sutherland law
where µ0 is a known viscosity at a known absolute temperature T0 (usually 273
K). The constants n and S are fit to the data, and both formulas are adequate over
a wide range of temperatures. For air, n ≈ 0.7 and S ≈ 110 K = 199°R. Other
values are given in Ref. 26.
Liquid viscosity decreases with temperature and is roughly exponential, µ ≈
ae−bT; but a better fit is the empirical result that ln µ is quadratic in 1/T, where
T is absolute temperature:
ln
μ
T0
T0 2
≈a+b( )+c( ) μ0
T
T
(1.28)
For water, with T0 = 273.16 K, µ0 = 0.001792 kg/(m · s), suggested values are
a = −1.94, b = −4.80, and c = 6.74, with accuracy about ±1 percent. The
viscosity of water is tabulated in Table A.1. For further viscosity data, see Refs.
27, 28, and 29.
Nonnewtonian Fluids
Fluids that do not follow the linear law of Eq. (1.23) are called nonnewtonian
and are treated in books on rheology [16]. Figure 1.9a compares some examples
to a newtonian fluid. For the nonlinear curves, the slope at any point is called
the apparent viscosity.
Dilatant. This fluid is shear-thickening, increasing its resistance with increasing
strain rate. Examples are suspensions of corn starch or sand in water. The classic
case is quicksand, which stiffens up if one thrashes about.
Pseudoplastic. A shear-thinning fluid is less resistant at higher strain rates. A
very strong thinning is called plastic. Some of the many examples are polymer
solutions, colloidal suspensions, paper pulp in water, latex paint, blood plasma,
32
Chapter 1 Introduction
Shear
stress
τ
Ideal Bingham
plastic
Dilatant
Plastic
Newtonian
Yield
stress
Shear
stress
τ
Rheopectic
Common
fluids
Pseudoplastic
Thixotropic
Constant
strain rate
0
Shear strain rate d θ
dt
(a)
0
Time
(b)
Fig. 1.9 Rheological behavior of various viscous materials: (a) stress versus strain rate;
(b) effect of time on applied stress.
syrup, and molasses. The classic case is paint, which is thick when poured but
thin when brushed at a high strain rate.
Bingham plastic. The limiting case of a plastic substance is one that requires a
finite yield stress before it begins to flow. Figure 1.9a shows yielding followed
by linear behavior, but nonlinear flow can also occur. Some examples are clay
suspensions, drilling mud, toothpaste, mayonnaise, chocolate, and mustard. The
classic case is catsup, which will not come out of the bottle until you stress it by
shaking.
A further complication of nonnewtonian behavior is the transient effect shown
in Fig. 1.9b. Some fluids require a gradually increasing shear stress to maintain
a constant strain rate and are called rheopectic. The opposite case of a fluid that
thins out with time and requires decreasing stress is termed thixotropic. We
neglect nonnewtonian effects in this book; see Ref. 16 for further study.
Rheometers
Fig. 1.10 A rotating parallel-disk
rheometer (NETZSCH).
There are many commercial devices for measuring the shear stress versus strain
rate behavior of both newtonian and nonnewtonian fluids. They are generically
called rheometers and have various designs: parallel disks, cone-plate, rotating
coaxial cylinders, torsion, extensional, and capillary tubes. Reference 29 gives a
good discussion. A popular device is the parallel-disk rheometer, shown in
Fig. 1.10. A thin layer of fluid is placed between the disks, one of which rotates.
The resisting torque on the rotating disk is proportional to the viscosity of the
fluid. A simplified theory for this device is given in Example 1.10.
1.7 Viscosity and Other Secondary Properties 33
0.080
Υ, N/m
0.070
0.060
0.050
Fig. 1.11 Surface tension of a clean
air–water interface. Data from
­Table A.5.
0
10
20
30
40
50
60
70
80
90
100
T, °C
Surface Tension
A liquid, being unable to expand freely, will form an interface with a second
liquid or gas. The physical chemistry of such interfacial surfaces is quite complex,
and whole textbooks are devoted to this specialty [30]. Molecules deep within
the liquid repel each other because of their close packing. Molecules at the surface
are less dense and attract each other. Since half of their neighbors are missing,
the mechanical effect is that the surface is in tension. We can account adequately
for surface effects in fluid mechanics with the concept of surface tension.
If a cut of length dL is made in an interfacial surface, equal and opposite forces
of magnitude Υ dL are exposed normal to the cut and parallel to the surface,
where Υ is called the coefficient of surface tension. The dimensions of Υ are
{F/L}, with SI units of newtons per meter and BG units of pounds-force per foot.
An alternate concept is to open up the cut to an area dA; this requires work to
be done of amount Υ dA. Thus the coefficient Υ can also be regarded as the
surface energy per unit area of the interface, in N · m/m2 or ft · lbf/ft2.
The two most common interfaces are water–air and mercury–air. For a clean
surface at 20°C = 68°F, the measured surface tension is
0.0050 lbf/ft = 0.073 N/m
Y={
0.033 lbf/ft = 0.48 N/m
air–water
air–mercury
(1.29)
These are design values and can change considerably if the surface contains
contaminants like detergents or slicks. Generally Υ decreases with liquid temperature and is zero at the critical point. Values of Υ for water are given in
Fig. 1.11 and Table A.5.
If the interface is curved, a mechanical balance shows that there is a pressure
difference across the interface, the pressure being higher on the concave side, as
illustrated in Fig. 1.12. In Fig. 1.12a, the pressure increase in the interior of a
liquid cylinder is balanced by two surface-tension forces:
2RL Δp = 2ΥL
or
Δp =
Υ
R
(1.30)
34
Chapter 1 Introduction
πR 2 ∆p
2RL ∆p
∆p dA
ΥL
ΥdL 1
2πRΥ
ΥdL 2
ΥL
R2
R1
L
ΥdL 2
ΥdL 1
2R
(a)
(b)
(c)
Fig. 1.12 Pressure change across a curved interface due to surface tension: (a) interior of a liquid cylinder; (b) interior of a spherical droplet; (c) general curved interface.
We are not considering the weight of the liquid in this calculation. In Fig. 1.12b,
the pressure increase in the interior of a spherical droplet balances a ring of
surface-tension force:
πR2 Δp = 2πRΥ
or
Δp =
2Υ
R
(1.31)
We can use this result to predict the pressure increase inside a soap bubble, which
has two interfaces with air, an inner and outer surface of nearly the same radius R:
Δpbubble ≈ 2 Δpdroplet =
4Υ
R
(1.32)
Figure 1.12c shows the general case of an arbitrarily curved interface whose
principal radii of curvature are R1 and R2. A force balance normal to the surface
will show that the pressure increase on the concave side is
Δp = Υ(R1−1 + R2−1 ) (1.33)
Equations (1.30) to (1.32) can all be derived from this general relation; for example, in Eq. (1.30), R1 = R and R2 = ∞.
A second important surface effect is the contact angle θ, which appears when
a liquid interface intersects with a solid surface, as in Fig. 1.13. The force balance would then involve both Υ and θ. If the contact angle is less than 90°, the
liquid is said to wet the solid; if θ > 90°, the liquid is termed nonwetting. For
example, water wets soap but does not wet wax. Water is extremely wetting to
a clean glass surface, with θ ≈ 0°. Like Υ, the contact angle θ is sensitive to
the actual physicochemical conditions of the solid–liquid interface. For a clean
mercury–air–glass interface, θ = 130°.
1.7 Viscosity and Other Secondary Properties 35
Gas
Liquid
Fig. 1.13 Contact-angle effects at
liquid–gas–solid interface. If
θ < 90°, the liquid “wets” the solid;
if θ > 90°, the liquid is nonwetting.
θ
Nonwetting
θ
Solid
Example 1.8 illustrates how surface tension causes a fluid interface to rise or
fall in a capillary tube.
EXAMPLE 1.8
Derive an expression for the change in height h in a circular tube of a liquid with
surface tension Υ and contact angle θ, as in Fig. E1.8.
θ
Solution
The vertical component of the ring surface-tension force at the interface in the tube
must ­balance the weight of the column of fluid of height h:
h
2πRΥ cos θ = γπR2h
Solving for h, we have the desired result:
2R
E1.8
h=
2Υ cos θ
γR
Ans.
Thus the capillary height increases inversely with tube radius R and is positive if
θ < 90° (­wetting liquid) and negative (capillary depression) if θ > 90°.
Suppose that R = 1 mm. Then the capillary rise for a water–air–glass interface,
θ ≈ 0°, Υ = 0.073 N/m, and ρ = 1000 kg/m3 is
h=
2(0.073 N/m) (cos 0°)
(1000 kg/m3 ) (9.81 m/s2 ) (0.001 m)
= 0.015 (N · s2 )/kg = 0.015 m = 1.5 cm
For a mercury–air–glass interface, with θ = 130°, Υ = 0.48 N/m, and ρ = 13,600 kg/m3,
the capillary rise is
h=
2(0.48) (cos 130°)
= −0.0046 m = −0.46 cm
13,600(9.81) (0.001)
When a small-diameter tube is used to make pressure measurements (Chap. 2), these
capillary effects must be corrected for.
36
Chapter 1 Introduction
Vapor Pressure
Vapor pressure is the pressure at which a liquid boils and is in equilibrium with
its own vapor. For example, the vapor pressure of water at 68°F is 49 lbf/ft2,
while that of mercury is only 0.0035 lbf/ft2. If the liquid pressure is greater than
the vapor pressure, the only exchange between liquid and vapor is evaporation at
the interface. If, however, the liquid pressure falls below the vapor pressure, vapor
bubbles begin to appear in the liquid. If water is heated to 212°F, its vapor pressure
rises to 2116 lbf/ft2, and thus water at normal atmospheric pressure will boil.
When the liquid pressure is dropped below the vapor pressure due to a flow
phenomenon, we call the process cavitation. If water is accelerated from rest to
about 50 ft/s, its pressure drops by about 15 lbf/in2, or 1 atm. This can cause
cavitation [31].
The dimensionless parameter describing flow-induced boiling is the cavitation
number
Ca =
where pa
pv
V
ρ
=
=
=
=
pa − pv
1
2
2 ρV
(1.34)
ambient pressure
vapor pressure
characteristic flow velocity
fluid density
Depending on the geometry, a given flow has a critical value of Ca below which
the flow will begin to cavitate. Values of surface tension and vapor pressure of
water are given in Table A.5. The vapor pressure of water is plotted in Fig. 1.14.
Figure 1.15a shows cavitation bubbles being formed on the low-pressure surfaces of a marine propeller, as happens in water being churned by the propeller.
When these bubbles move into a higher-pressure region, they collapse implosively. Cavitation collapse can rapidly spall and erode metallic surfaces and eventually destroy them, as shown in Fig. 1.15b.
100
pv, kPa
80
60
40
20
Fig. 1.14 Vapor pressure of water.
Data from Table A.5.
0
0
20
40
T, °C
60
80
100
1.7 Viscosity and Other Secondary Properties 37
(a)
Fig. 1.15 Two aspects of cavitation
bubble formation in liquid flows:
(a) Beauty: spiral bubble sheets
form from the surface of a marine
propeller (National Physical
­Laboratory/Crown Copyright/
Science Source); (b) ugliness:
­collapsing bubbles erode a propeller
surface (Flafabri/Alamy Stock
Photo).
(b)
38
Chapter 1 Introduction
EXAMPLE 1.9
A certain torpedo, moving in fresh water at 10°C, has a minimum-pressure point given
by the formula
pmin = p0 − 0.35 ρV2(1)
where p0 = 115 kPa, ρ is the water density, and V is the torpedo velocity. Estimate the
velocity at which cavitation bubbles will form on the torpedo. The constant 0.35 is
­dimensionless.
Solution
∙ Assumption: Cavitation bubbles form when the minimum pressure equals the vapor
­pressure pv.
∙ Approach: Solve Eq. (1) above, which is related to the Bernoulli equation from
­Example 1.3, for the velocity when pmin = pv. Use SI units (m, N, kg, s).
∙ Property values: At 10°C, read Table A.1 for ρ = 1000 kg/m3 and Table A.5 for
pv = 1.227 kPa.
∙ Solution steps: Insert the known data into Eq. (1) and solve for the velocity, using
SI units:
pmin = pv = 1227 Pa = 115,000 Pa − 0.35 (1000
Solve V2 =
m3 )
(115,000 − 1227)
m2
= 325 2 or V =
0.35(1000)
s
kg
V2, with V in m/s
√325
≈ 18.0 m/s
Ans.
∙ Comments: Note that the use of SI units requires no conversion factors, as discussed
in Example 1.3b. Pressures must be entered in pascals, not kilopascals.
No-Slip and No-Temperature-Jump Conditions
When a fluid flow is bounded by a solid surface, molecular interactions cause
the fluid in contact with the surface to seek momentum and energy equilibrium
with that surface. All liquids essentially are in equilibrium with the surfaces they
contact. All gases are, too, except under the most rarefied conditions [18]. Excluding rarefied gases, then, all fluids at a point of contact with a solid take on the
velocity and temperature of that surface:
Vfluid ≡ Vwall
Tfluid ≡ Twall
(1.35)
These are called the no-slip and no-temperature-jump conditions, respectively.
They serve as boundary conditions for analysis of fluid flow past a solid surface.
­Figure 1.16 shows the velocity profiles for a fluid flow over a solid surface with
no-slip and slip conditions. The no-slip boundary condition assumes that the
velocity of the fluid layer in direct contact with the boundary is identical to the
velocity of this boundary (Eq. 1.35). There is no relative movement between the
boundary and this fluid layer; therefore there is no slip. On the other hand, the
slip boundary condition assumes a discontinuity in the velocity function that
results in a relative movement between the fluid and the boundary; therefore there
is slip. The hypothetical vertical distance inside of the boundary at which the
velocity function would effectively reach the velocity of the boundary is referred
to as slip length l.
1.7 Viscosity and Other Secondary Properties 39
No-slip condition
y
u
y
Slip condition
u
δuwall = 0
δuwall ≈ l
δuwall
Fig. 1.16 Schematics of velocity
profiles over a solid surface with
no-slip and slip conditions.
∂u
∂y
wall
l
In some applications, the no-slip condition is partially relaxed in the analysis
of inviscid flow (Chap. 8). The flow is allowed to “slip” past the surface but not
to permeate through the surface
Vnormal (fluid) ≡ Vnormal (solid)
(1.36)
while the tangential velocity Vt is allowed to be independent of the wall. The
analysis is much simpler, but the flow patterns are highly idealized.
For high-viscosity newtonian fluids, the linear velocity assumption and the
no-slip conditions can yield some sophisticated approximate analyses for two- and
three-dimensional viscous flows. The following example, for a type of rotatingdisk viscometer, will illustrate.
EXAMPLE 1.10
An oil film of viscosity µ and thickness h ≪ R lies between a solid wall and a circular disk, as in Fig. E1.10. The disk is rotated steadily at angular velocity Ω. Noting
that both velocity and shear stress vary with radius r, derive a formula for the torque
M required to rotate the disk. Neglect air drag.
Solution
∙ System sketch: Figure E1.10 shows a side view (a) and a top view (b) of the system.
∙ Assumptions: Linear velocity profile, laminar flow, no-slip, local shear stress given
by Eq. (1.23).
Oil film
thickness
h
Ω
r=R
r=R
dM = (τdA)r
r
r
dA = 2πr dr
Fixed wall
dr
E1.10
(a)
(b)
40
Chapter 1 Introduction
∙ Approach: Estimate the shear stress on a circular strip of width dr and area dA = 2πr
dr in Fig. E1.10b, then find the moment dM about the origin caused by this shear stress.
Integrate over the entire disk to find the total moment M.
∙ Property values: Constant oil viscosity µ. In this steady flow, oil density is not relevant.
∙ Solution steps: At radius r, the velocity in the oil is tangential, varying from zero
at the fixed wall (no-slip) to u = Ωr at the disk surface (also no-slip). The shear
stress at this position is thus
τ=μ
du
Ωr
≈μ
dy
h
This shear stress is everywhere perpendicular to the radius from the origin (see Fig. E1.10b).
Then the total moment about the disk origin, caused by shearing this circular strip, can be
found and integrated:
μΩr
2πμΩ
dM = (τ) (dA)r = (
(2πr dr)r, M = dM =
h )
h
∫
R
4
∫ r dr = πμΩR
2h
3
Ans.
0
∙ Comments: This is a simplified engineering analysis, which neglects possible edge
effects, air drag on the top of the disk, and the turbulence that might ensue if the
disk rotates too fast.
Slip Flow in Rarefied Gases
The “free slip” boundary condition, Eq. (1.36), is an unrealistic mathematical artifice to enable inviscid-flow solutions. However, actual, realistic wall slip occurs
in rarefied gases as shown in Fig. 1.16, where there are too few molecules to
establish momentum equilibrium with the wall. In 1879, the physicist James Clerk
Maxwell used the kinetic theory of gases to predict a slip velocity at the wall:
δuwall ≈ ℓ
∂u
|wall ∂y
(1.37)
where ℓ is the mean free path of the gas, and u and x are along the wall. If ℓ is very
small compared to the lateral scale L of the flow, the Knudsen number, Kn = ℓ/L,
is small, and the slip velocity is near zero. We will assign a few slip problems, but
the details of rarefied gas flow are left for further reading in Refs. 18 and 46.
Speed of Sound
In gas flow, one must be aware of compressibility effects (significant density
changes caused by the flow). We shall see in Sec. 4.2 and in Chap. 9 that compressibility becomes important when the flow velocity reaches a significant fraction of the speed of sound of the fluid. The speed of sound a of a fluid is the
rate of propagation of small-disturbance pressure pulses (“sound waves”) through
the fluid. In Chap. 9 we shall show, from momentum and thermodynamic arguments, that the speed of sound is defined by a pressure-density derivative proportional to the isentropic bulk modulus:
a2 =
cp
β
∂p
∂p
=
= k( ) , k =
(
)
ρ
cv
∂ρ s
∂ρ T
1.8 Flow Patterns: Streamlines, Pathlines, and Streaklines 41
∂p
where β = isentropic bulk modulus = ρ ( ) .
∂ρ s
This is true for either a liquid or a gas, but it is for gases that the problem of
compressibility occurs. For an ideal gas, Eq. (1.10), we obtain the simple formula
aideal gas = (kRT) 1/2
(1.38)
where R is the gas constant, Eq. (1.11), and T the absolute temperature. For
example, for air at 20°C, a = {(1.40)[287 m2/(s2 · K)](293 K)}1/2 ≈ 343 m/s (1126
ft/s = 768 mi/h). If, in this case, the air velocity reaches a significant fraction of
a, say, 100 m/s, then we must account for compressibility effects (Chap. 9).
Another way to state this is to account for compressibility when the Mach number, Ma = V/a, of the flow reaches about 0.3.
The speed of sound of water is tabulated in Table A.5. For near perfect gases,
like air, the speed of sound is simply calculated by Eq. (1.38). Many liquids have
their bulk modulus listed in Table A.3. Note, however, as discussed in Ref. 45,
even a very small amount of dissolved gas in a liquid can reduce the mixture
speed of sound by up to 80 percent.
EXAMPLE 1.11
A commercial airplane flies at 540 mi/h at a standard altitude of 30,000 ft. What is its
Mach number?
Solution
∙ Approach: Find the “standard” speed of sound; divide it into the velocity, using
proper units.
∙ Property values: From Table A.6, at 30,000 ft (9144 m), a ≈ 303 m/s. Check this
against the standard temperature, estimated from the table to be 229 K. From
Eq. (1.38) for air,
a = [kRairT] 1/2 = [1.4(287) (229) ] 1/2 ≈ 303 m/s.
∙ Solution steps: Convert the airplane velocity to m/s:
V = (540 mi/h) [0.44704 m/s/(mi/h) ] ≈ 241 m/s.
Then the Mach number is given by
Ma = V/a = (241 m/s)/(303 m/s) = 0.80
Ans.
∙ Comments: This value, Ma = 0.80, is typical of present-day commercial airliners.
1.8 Flow Patterns: Streamlines, Pathlines, and Streaklines
Fluid mechanics is a highly visual subject. The patterns of flow can be visualized
in a dozen different ways, and you can view these sketches or photographs and
learn a great deal qualitatively and often quantitatively about the flow.
Four basic types of line patterns are used to visualize flows:
1. A streamline is a line everywhere tangent to the velocity vector at a given instant.
2. A pathline is the actual path traversed by a given fluid particle.
42
Chapter 1 Introduction
V
No flow across
streamtube walls
Fig. 1.17 The most common method
of flow-pattern presentation:
(a) Streamlines are everywhere
­tangent to the local velocity vector;
(b) a streamtube is formed by a
closed collection of streamlines.
Individual
streamline
(a)
(b)
3. A streakline is the locus of particles that have earlier passed through a prescribed point [33].
4. A timeline is a set of fluid particles that form a line at a given instant.
The streamline is convenient to calculate mathematically, while the other three are
easier to generate experimentally. Note that a streamline and a timeline are instantaneous lines, while the pathline and the streakline are generated by the passage of time.
A pathline can be found by a time exposure of a single marked particle moving
through the flow. Streamlines are difficult to generate experimentally in unsteady flow
unless one marks a great many particles and notes their direction of motion during a
very short time interval [32]. In steady flow, where velocity varies only with position,
the situation simplifies greatly:
Streamlines, pathlines, and streaklines are identical in steady flow.
In fluid mechanics the most common mathematical result for visualization
purposes is the streamline pattern. Figure 1.17a shows a typical set of streamlines,
and Fig. 1.17b shows a closed pattern called a streamtube. By definition the fluid
within a streamtube is confined there because it cannot cross the streamlines; thus
the streamtube walls need not be solid but may be fluid surfaces.
Figure 1.18 shows an arbitrary velocity vector. If the elemental arc length dr of
a streamline is to be parallel to V, their respective components must be in proportion:
Streamline:
dx dy dz dr
=
=
= u
v
w
V
(1.39)
If the velocities (u, v, w) are known functions of position and time, Eq. (1.39)
can be integrated to find the streamline passing through the initial point (x0,
y0, z0, t0). The method is straightforward for steady flows but may be laborious for unsteady flow.
Flow Visualization
Clever experimentation can produce revealing images of a fluid flow pattern, as
shown earlier in Figs. 1.15a and 1.16. For example, streaklines are produced by
the continuous release of marked particles (dye, smoke, or bubbles) from a given
1.8 Flow Patterns: Streamlines, Pathlines, and Streaklines 43
z
V
V
dr
dz
dx
u
w
y
dy
v
Fig. 1.18 Geometric relations for
defining a streamline.
x
point. If the flow is steady, the streaklines will be identical to the streamlines and
pathlines of the flow.
Some methods of flow visualization include the following [34–38]:
1.
2.
3.
4.
5.
6.
7.
8.
Dye, smoke, or bubble discharges.
Surface powder or flakes on liquid flows.
Floating or neutral-density particles.
Optical techniques that detect density changes in gas flows: shadowgraph,
schlieren, and interferometer.
Tufts of yarn attached to boundary surfaces.
Evaporative coatings on boundary surfaces.
Luminescent fluids, additives, or bioluminescence.
Particle image velocimetry (PIV).
Figure 1.15a was visualized by bubble releases. Figure 1.19 illustrates the huge
size of the wingtip vortices from the wing of an agricultural plane [42]. The
visualization is made by a technique that uses colored smoke rising from the
ground. The swirl at the wingtip traces the aircraft’s wake vortices, which exerts
a powerful influence on the flow field behind the plane. The wingtip vortices can
pose a hazard to aircraft, especially during the landing and takeoff phases of
flight. NASA researchers studied wake vortices in actual flight tests in research
aircraft. Their goal is to fully understand the phenomenon, then use that knowledge to create an automated system that could predict changing wake vortex
conditions at airports.
Mathematical details of streamline/streakline/pathline analysis are given in
Ref. 33. References 39–41 are beautiful albums of flow photographs. References
34–36 are monographs on flow visualization techniques.
Fluid mechanics is a marvelous subject for visualization, not just for still
(steady) patterns, but also for moving (unsteady) motion studies. An outstanding
list of available flow movies and videotapes is given by Carr and Young [43].
44
Chapter 1 Introduction
Fig. 1.19 Visualization of wingtip
vortices, illustrating the huge size of
the vortices produced during takeoff
phase of an airplane. (NASA)
1.9 Basic Flow Analysis Techniques
Forces and motion are the main focuses in both solid and fluid mechanics. For
solid mechanics, we have used the free-body diagram extensively for the analysis.
But for fluid mechanics, we will use a system or a control volume, depending on
the problem being analyzed. The relations between the forces and motion are
governed by conversion of mass and Newton’s second law of motion. However,
the expressions of these basic laws in a system or a control volume is different.
Integral and Differential Approaches
In analyzing fluid motion with a control volume, we might take one of two
approaches: (1) formulating the basic laws of fluid mechanics in a finite region,
making a balance of flow in versus flow out, and determining gross flow effects
such as the force or torque on a body or the total energy exchange; or (2) formulating the basic laws of fluid mechanics in an infinitesimal control volume, and seeking to describe the detailed flow pattern at every point (x, y, z) in the flow fluid.
In the first case the resulting equations are integral formulations of the basic
laws introduced in detail in Chap. 3. An example of applications might be the
overall anchoring force required for a section of pipes or the mass flow rate for
an open jet. The second case is the “differential” approach and is developed in
Chap. 4. The resulting equations are differential equations. Solution of the differential equations of fluid motion provides a means of determining the detailed
behavior of the flow field. An example might be the pressure distribution on an
1.10 The Fundamentals of Engineering (FE) Examination 45
airplane wing surface or the velocity distribution in a ventilated clean room.
These approaches, together with the experimental study based on the dimensional
analysis and similarity introduced in Chap. 5, construct the basic flow analysis
techniques for fluid mechanics and provide necessary means to solve different
types of complex fluid dynamics problems.
It should be noted that with the availability of powerful computers it is feasible to solve the differential equations using the techniques of numerical analysis known as CFD. Although it is beyond the scope of this book to look extensively
into this approach, the reader should be aware that CFD has become a common
engineering tool for a large class of fluid flow, heat, and mass transfer problems.
Basic Laws Governing the Fluid Motion
In all cases, the flow must satisfy the three basic laws of mechanics plus a thermodynamic state relation and associated boundary conditions:
1.
2.
3.
4.
5.
Conservation of mass (continuity).
Linear momentum (Newton’s second law).
First law of thermodynamics (conservation of energy).
A state relation like ρ = ρ(p, T ).
Appropriate boundary conditions at solid surfaces, interfaces, inlets, and exits.
In integral and differential analyses, these five relations are modeled mathematically and solved by computational methods. In an experimental study, the
fluid itself performs this task without the use of any mathematics. In other
words, these laws are believed to be fundamental to physics, and no fluid flow
is known to violate them.
1.10 The Fundamentals of Engineering (FE) Examination
The road toward a professional engineer’s license has a first stop, the Fundamentals of Engineering Examination, known as the FE exam. It was formerly known
as the Engineer-in-Training (E-I-T) Examination. This eight-hour national test will
probably soon be required of all engineering graduates, not just for licensure, but
as a student assessment tool. The 120-problem, four-hour morning session covers
many general studies:
Mathematics—15%
Ethics and business
Material properties—7%
practices—7%
Engineering probability
Engineering
Fluid mechanics—7%
and statistics—7% economics—8%
Chemistry—9%
Engineering
Electricity and
mechanics—10% magnetism—9%
Computers—7%
Strength of
Thermodynamics—7%
materials—7%
For the 60-problem, four-hour afternoon session you may choose one of seven
modules: chemical, civil, electrical, environmental, industrial, mechanical, and
46
Chapter 1 Introduction
other/­general engineering. Note that fluid mechanics is an integral topic of the
examination. Therefore, for practice, this text includes a number of end-of-­chapter
FE problems where appropriate.
The format for the FE exam questions is multiple-choice, usually with five
selections, chosen carefully to tempt you with plausible answers if you used
incorrect units, forgot to double or halve something, are missing a factor of π, or
the like. In some cases, the selections are unintentionally ambiguous, such as the
following example from a previous exam:
Transition from laminar to turbulent flow occurs at a Reynolds number of
(A) 900 (B) 1200 (C) 1500 (D) 2100 (E) 3000
The “correct” answer was graded as (D), Re = 2100. Clearly the examiner was
thinking, but forgot to specify, Red for flow in a smooth circular pipe, since (see
Chaps. 6 and 7) transition is highly dependent on geometry, surface roughness,
and the length scale used in the definition of Re. The moral is not to get peevish
about the exam but simply to go with the flow (pun intended) and decide which
answer best fits an undergraduate training situation. Every effort has been made
to keep the FE exam questions in this text unambiguous.
1.11 The History of Fluid Mechanics
Many distinguished workers have contributed to the development of fluid mechanics. If you are a student, however, this is probably not the time to be studying
history. Later, during your career, you will enjoy reading about the history of, not
just fluid mechanics, but all of science. Here are some names that will be
­mentioned as we encounter their contributions in the rest of this book.
Name
Important Contribution
Archimedes (285–212 BC)
Established laws of buoyancy and floating bodies.
Leonardo da Vinci (1452–1519)
Formulated the first equation of continuity.
Isaac Newton (1642–1727)
Postulated the law of linear viscous stresses.
Leonhard Euler (1707–1783)Developed Bernoulli’s equation by solving the basic
equations.
L. M. H. Navier (1785–1836)Formulated the basic differential equations of
viscous flow.
Jean Louis Poiseuille (1799–1869)
Performed first experiments on laminar flow in tubes.
Osborne Reynolds (1842–1912)
Explained the phenomenon of transition to turbulence.
Ludwig Prandtl (1875–1953)Formulated boundary-layer theory, predicting flow
separation.
Theodore von Kármán (1881–1963)
Major advances in aerodynamics and turbulence theory.
References 12–15 provide a comprehensive treatment of the history of fluid
mechanics.
Summary
This chapter has discussed the behavior of a fluid—which, unlike a solid, must move
if subjected to a shear stress—and the important fluid properties. The writer believes
the most important property to be the velocity vector field V(x, y, z, t). Following
Problems 47
closely are the pressure p, density ρ, and temperature T. Many secondary properties
enter into various flow problems: viscosity µ, thermal conductivity k, specific weight
γ, surface tension Υ, speed of sound a, and vapor pressure pv. You must learn to
locate and use all these properties to become proficient in fluid mechanics.
There was a brief discussion of the five different kinds of mathematical relations we will use to solve flow problems—mass conservation, linear momentum,
first law of thermodynamics, equations of state, and appropriate boundary conditions at walls and other boundaries.
Flow patterns are also discussed briefly. The most popular, and useful, scheme
is to plot the field of streamlines, that is, lines everywhere parallel to the local
velocity vector.
Since the earth is 75 percent covered with water and 100 percent covered with
air, the scope of fluid mechanics is vast and touches nearly every human endeavor.
The sciences of meteorology, physical oceanography, and hydrology are concerned
with naturally occurring fluid flows, as are medical studies of breathing and blood
circulation. All transportation problems involve fluid motion, with well-developed
specialties in aerodynamics of aircraft and rockets and in naval hydrodynamics of
ships and submarines. Almost all our electric energy is developed either from
water flow, air flow through wind turbines, or from steam flow through turbine
generators. All combustion problems involve fluid motion as do the more classic
problems of irrigation, flood control, water supply, sewage disposal, projectile
motion, and oil and gas pipelines. The aim of this book is to present enough
fundamental concepts and practical applications in fluid mechanics to prepare you
to move smoothly into any of these specialized fields of the science of flow—and
then be prepared to move out again as new technologies develop.
Problems
Most of the problems herein are fairly straightforward. More
difficult or open-ended assignments are labeled with an asterisk as in Prob. 1.18. Problems labeled with a computer icon
may require the use of a computer. The standard end-ofchapter problems 1.1 to 1.85 (categorized in the problem list
below) are followed by fundamentals of engineering (FE)
exam problems FE1.1 to FE1.10 and comprehensive problems
C1.1 to C1.12.
The concept of a fluid
P1.1
P1.2
Problem Distribution
Section
1.1–1.3
1.4
1.6
1.7
1.7
1.7
1.7
1.8
1.11
Topic
Fluid continuum concept
Dimensions and units
Thermodynamic properties
Viscosity, no-slip condition
Surface tension
Vapor pressure; cavitation
Speed of sound, Mach number
Streamlines
History of fluid mechanics
Problems
1.1–1.4
1.5–1.23
1.24–1.37
1.38–1.61
1.62–1.71
1.72–1.74
1.75–1.80
1.81–1.83
1.84–1.85a–n
P1.3
A gas at 20°C may be considered rarefied, deviating
from the continuum concept, when it contains less than
1012 molecules per cubic millimeter. If Avogadro’s number is 6.023 × 1023 molecules per mole, what absolute
pressure (in Pa) for air does this r­epresent?
Table A.6 lists the density of the standard atmosphere as
a function of altitude. Use these values to estimate,
crudely—say, within a factor of 2—the number of molecules of air in the entire atmosphere of the earth.
For the triangular element in Fig. P1.3, show that a tilted free
liquid surface, in contact with an atmosphere at pressure pa,
must undergo shear stress and hence begin to flow. Hint:
Account for the weight of the fluid and show that a no-shear
condition will cause horizontal forces to be out of balance.
pa
θ
P1.3
Fluid density ρ
48
P1.4
Chapter 1 Introduction
Sand, and other granular materials, appear to flow; that
is, you can pour them from a container or a hopper.
There are whole textbooks on the “transport” of granular
materials [48]. Therefore, is sand a fluid? Explain.
Dimensions and units
P1.5
The mean free path of a gas, l, is defined as the average
distance traveled by molecules between collisions. A
proposed formula for estimating l of an ideal gas is
l = 1.26
P1.6
P1.11 In English Engineering units, the specific heat cp of
air at room temperature is approximately 0.24 Btu/
(lbm-°F). When working with kinetic energy relations, it is more appropriate to express cp as a velocity-squared per absolute degree. Express the numerical
value, in this form, of cp for air in (a) SI units, and (b)
BG units.
P1.12 For low-speed (laminar) steady flow through a circular
pipe, as shown in Fig. P1.12, the velocity u varies with
­radius and takes the form
μ
u=B
ρ √RT
What are the dimensions of the constant 1.26? Use the
formula to estimate the mean free path of air at 20°C and
7 kPa. Would you consider air rarefied at this ­condition?
Henri Darcy, a French engineer, proposed that the pressure drop Δp for flow at velocity V through a tube of
length L could be correlated in the form
Δp
= α LV2
ρ
Δp 2
(r − r 2 )
μ 0
where µ is the fluid viscosity and Δp is the pressure drop
from entrance to exit. What are the dimensions of the
constant B?
Pipe wall
r
r = r0
u (r)
If Darcy’s formulation is consistent, what are the dimensions of the coefficient α?
r=0
P1.7 Convert the following inappropriate quantities into SI
units: (a) 2.283 × 107 U.S. gallons per day; (b) 4.5
­furlongs per minute (racehorse speed); and (c) 72,800
avoirdupois ounces per acre.
P1.8 Suppose we know little about the strength of materials
but are told that the bending stress σ in a beam is P1.12
­proportional to the beam half-thickness y and also depends on the bending moment M and the beam area P1.13 The efficiency η of a pump is defined as the (dimensionmoment of inertia I. We also learn that, for the parless) ratio of the power developed by the flow to the
ticular case M = 2900 in · lbf, y = 1.5 in, and I = 0.4
power required to drive the pump:
in4, the predicted stress is 75 MPa. Using this inforQΔp
mation and dimensional reasoning only, find, to three
η=
significant figures, the only possible dimensionally
input power
homogeneous formula σ = y f (M, I ).
P1.9 A hemispherical container, 26 inches in diameter, is
where Q is the volume rate of flow and Δp is the presfilled with a liquid at 20°C and weighed. The liquid
sure rise produced by the pump. Suppose that a certain
weight is found to be 1617 ounces. (a) What is the denpump develops a pressure rise of 35 lbf/in2 when its flow
sity of the fluid, in kg/m3? (b) What fluid might this be?
rate is 40 L/s. If the input power is 16 hp, what is the
Assume standard gravity, g = 9.807 m/s2.
­efficiency?
P1.10 The Stokes–Oseen formula [33] for drag force F on a *P1.14 Figure P1.14 shows the flow of water over a dam. The
sphere of diameter D in a fluid stream of low velocity V,
volume flow Q is known to depend only on crest width
density ρ, and viscosity µ is
B, acceleration of gravity g, and upstream water height
H above the dam crest. It is further known that Q is
9π 2 2
F = 3πμDV + ρV D
proportional to B. What is the form of the only possi16
ble dimensionally homogeneous relation for this flow
rate?
Is this formula dimensionally homogeneous?
Problems 49
Water level
Q
H
Dam
B
P1.14 P1.15 The height H that fluid rises in a liquid barometer tube
­depends upon the liquid density ρ, the barometric pressure p, and the acceleration of gravity g. (a) Arrange
these four variables into a single dimensionless group.
(b) Can you deduce (or guess) the numerical value of
your group?
P1.16 Algebraic equations such as Bernoulli’s relation, Eq. (1)
of Example 1.3, are dimensionally consistent, but what
about differential equations? Consider, for example, the
boundary-layer x-momentum equation, first derived by
Ludwig Prandtl in 1904:
ρu
∂p
∂u
∂u
∂τ
+ ρν = − + ρgx +
∂x
∂y
∂x
∂y
where τ is the boundary-layer shear stress and gx is the
component of gravity in the x direction. Is this equation
dimensionally consistent? Can you draw a general conclusion?
P1.17 The Hazen–Williams hydraulics formula for volume rate
of flow Q through a pipe of diameter D and length L is
given by
Q ≈ 61.9 D2.63 (
Δp 0.54
L )
where Δp is the pressure drop required to drive the flow.
What are the dimensions of the constant 61.9? Can this
formula be used with confidence for various liquids and
gases?
*P1.18 For small particles at low velocities, the first term in the
Stokes–Oseen drag law, Prob. 1.10, is dominant; hence,
F ≈ KV, where K is a constant. Suppose a particle of
mass m is constrained to move horizontally from the initial position x = 0 with initial velocity V0. Show (a) that
its velocity will decrease exponentially with time and (b)
that it will stop after traveling a distance x = mV0/K.
P1.19 In his study of the circular hydraulic jump formed by a
faucet flowing into a sink, Watson [47] proposed a parameter combining volume flow rate Q, density ρ, and
viscosity µ of the fluid, and depth h of the water in the
sink. He claims that his grouping is dimensionless, with
Q in the numerator. Can you verify this?
P1.20 Books on porous media and atomization claim that the
viscosity µ and surface tension Υ of a fluid can be combined with a characteristic velocity U to form an important dimensionless parameter. (a) Verify that this is so.
(b) Evaluate this parameter for water at 20°C and a velocity of 3.5 cm/s. Note: You get extra credit if you know
the name of this parameter.
P1.21 Aeronautical engineers measure the pitching moment M0
of a wing and then write it in the following form for use
in other cases:
M0 = βV2 AC ρ
where V is the wing velocity, A the wing area, C the
wing chord length, and ρ the air density. What are the
dimensions of the coefficient β?
P1.22 The Ekman number, Ek, arises in geophysical fluid
­dynamics. It is a dimensionless parameter combining
seawater density ρ, a characteristic length L, seawater
viscosity µ, and the Coriolis frequency Ω sinφ, where
Ω is the rotation rate of the earth and φ is the latitude
angle. Determine the correct form of Ek if the viscosity
is in the numerator.
P1.23 During World War II, Sir Geoffrey Taylor, a British fluid
dynamicist, used dimensional analysis to estimate the
­energy released by an atomic bomb explosion. He assumed that the energy released, E, was a function of
blast wave radius R, air density ρ, and time t. Arrange
these variables into a single dimensionless group, which
we may term the blast wave number.
Thermodynamic properties
P1.24 Air, assumed to be an ideal gas with k = 1.40, flows isentropically through a nozzle. At section 1, conditions
are sea level standard (see Table A.6). At section 2, the
temperature is −50°C. Estimate (a) the pressure and (b)
the density of the air at section 2.
P1.25 On a summer day in Narragansett, Rhode Island, the air
temperature is 74°F and the barometric pressure is 14.5
lbf/in2. Estimate the air density in kg/m3.
P1.26 When we in the United States say a car’s tire is filled “to
32 lb,” we mean that its internal pressure is 32 lbf/in2
above the ambient atmosphere. If the tire is at sea level,
has a volume of 3.0 ft3, and is at 75°F, estimate the total
weight of air, in lbf, inside the tire.
50
Chapter 1 Introduction
P1.27 For steam at a pressure of 45 atm, some values of temperature and specific volume are as follows, from Ref. 23:
T, °F
v, ft3/lbm
500
600
700
800
900
0.7014
0.8464
0.9653
1.074
1.177
Find an average value of the predicted gas constant R in
m2/(s2 · K). Do these data reasonably approximate an
ideal gas? If not, explain.
P1.28 Wet atmospheric air at 100 percent relative humidity
contains saturated water vapor and, by Dalton’s law of
partial pressures,
patm = pdry air + pwater vapor
P1.29
P1.30
P1.31
P1.32
P1.33
P1.34
P1.35
P1.36
Suppose this wet atmosphere is at 40°C and 1 atm. Calculate the density of this 100 percent humid air, and compare it with the density of dry air at the same conditions.
A compressed-air tank holds 5 ft3 of air at 120 lbf/in2
“gage,” that is, above atmospheric pressure. Estimate the
work input, in ft-lbf, required to compress this air from
the atmosphere, assuming a reversible isothermal
­process.
Repeat Prob. 1.29 if the tank is filled with compressed
­water instead of air. Why is the work input thousands of
times less than that in Prob. 1.29?
One cubic foot of argon gas at 10°C and 1 atm is compressed isentropically to a pressure of 600 kPa. (a) What
will be its new pressure and temperature? (b) If it is
­allowed to cool at this new volume back to 10°C, what
will be the final pressure?
A blimp is approximated by a prolate spheroid 90 m long
and 30 m in diameter. Estimate the weight of 20°C gas
within the blimp for (a) helium at 1.1 atm and (b) air at
1.0 atm. What might the difference between these two
­values represent (see Chap. 2)?
A tank contains 9 kg of CO2 at 20°C and 2.0 MPa. Estimate the volume of the tank, in m3.
Consider steam at the following state near the saturation
line: (p1, T1) = (1.31 MPa, 290°C). Using the ideal gas
relation and the steam tables, respectively, calculate and
compare (a) the density ρ1 and (b) the density ρ2 if the
steam expands isentropically to a new pressure of 414
kPa. Discuss your results.
In Table A.4, most common gases (air, nitrogen, oxygen,
hydrogen) have a specific-heat ratio k ≈ 1.40. Why do
­argon and helium have such high values? Why does NH3
have such a low value? What is the lowest k for any gas
that you know of?
Experimental data [49] for the density of n-pentane
liquid for high pressures, at 50°C, are listed as
follows:
Pressure, kPa
3
Density, kg/m
100
10,230
20,700
34,310
586.3
604.1
617.8
632.8
(a) Fit these data to reasonably accurate values of B and
n from Eq. (1.19). (b) Evaluate ρ at 30 MPa.
P1.37 A near-ideal gas has a molecular weight of 44 and a
specific heat cv = 610 J/(kg · K). What are (a) its
­
specific-heat ratio, k, and (b) its speed of sound at 100°C?
Viscosity, no-slip condition
P1.38 In Fig. 1.8, if the fluid is glycerin at 20°C and the width
between plates is 6 mm, what shear stress (in Pa) is
­required to move the upper plate at 5.5 m/s? What is the
Reynolds number if L is taken to be the distance between
plates?
P1.39 Knowing µ for air at 20°C from Table 1.4, estimate its
viscosity at 500°C by (a) the power law and (b) the
Sutherland law. Also make an estimate from (c) Fig. 1.7.
Compare with the accepted value of µ ≈ 3.58 E−5
kg/m · s.
P1.40 Glycerin at 20°C fills the space between a hollow sleeve
of diameter 12 cm and a fixed coaxial solid rod of diameter 11.8 cm. The outer sleeve is rotated at 120 rev/min.
­Assuming no temperature change, estimate the torque
required, in N · m per meter of rod length, to hold the
inner rod fixed.
P1.41 An aluminum cylinder weighing 30 N, 6 cm in diameter
and 40 cm long, is falling concentrically through a long
vertical sleeve of diameter 6.04 cm. The clearance is
filled with SAE 50 oil at 20°C. Estimate the terminal
(zero ­acceleration) fall velocity. Neglect air drag and assume a linear velocity distribution in the oil. Hint: You
are given diameters, not radii.
P1.42 Helium at 20°C has a viscosity of 1.97 E−5 kg/(m · s).
Use the power law approximation and the data of Table
A.4 to estimate the temperature, in °C, at which helium’s
viscosity will double.
P1.43 For the flow of gas between two parallel plates of
Fig. 1.8, reanalyze for the case of slip flow at both walls.
Use the simple slip condition, δuwall = ℓ (du/dy)wall,
where ℓ is the mean free path of the fluid. Sketch the
expected velocity profile and find an expression for the
shear stress at each wall.
P1.44 One type of viscometer is simply a long capillary tube. A
commercial device is shown in Prob. C1.10. One measures the volume flow rate Q and the pressure drop Δp
and, of course, the radius and length of the tube. The
theoretical formula, which will be discussed in Chap. 6,
is Δp ≈ 8 μQL /(πR4 ). For a capillary of diameter 4 mm
and length 254 mm, the test fluid flows at 0.9 m3/h when
Problems 51
the pressure drop is 400 kPa. Find the predicted viscosity
in kg/m · s.
P1.45 A block of weight W slides down an inclined plane while
lubricated by a thin film of oil, as in Fig. P1.45. The film
contact area is A and its thickness is h. Assuming a linear
velocity distribution in the film, derive an expression for
the “terminal” (zero-acceleration) velocity V of the
block. Find the terminal velocity of the block if the block
mass is 6 kg, A = 35 cm2, θ = 15°, and the film is 1-mmthick SAE 30 oil at 20°C.
derive the force F required to pull the plate at velocity V.
(b) Is there a necessary relation between the two viscosities, µ1 and µ2?
h1
µ1
F, V
h2
µ2
Liquid film of
thickness h
P1.48
W
V
Block contact
area A
θ
P1.45 P1.46 A simple and popular model for two nonnewtonian fluids in Fig. 1.9a is the power law:
τ ≈ C(
du n
dy )
where C and n are constants fit to the fluid [16]. From
Fig. 1.9a, deduce the values of the exponent n for
which the fluid is (a) newtonian, (b) dilatant, and (c)
pseudoplastic. Consider the specific model constant
C = 0.4 N · sn/m2, with the fluid being sheared between two parallel plates as in Fig. 1.8. If the shear
stress in the fluid is 1200 Pa, find the velocity V of the
upper plate for the cases (d) n = 1.0, (e) n = 1.2, and
(f ) n = 0.8.
P1.47 Data for the apparent viscosity of average human blood,
at normal body temperature of 37°C, vary with shear
strain rate, as shown in the following table.
Strain rate, s−1
Apparent viscosity,
kg/(m · s)
1
10
100
1000
0.011
0.009
0.006
0.004
(a) Is blood a nonnewtonian fluid? (b) If so, what type of
fluid is it? (c) How do these viscosities compare with
plain water at 37°C?
P1.48 A thin plate is separated from two fixed plates by very
­viscous liquids µ1 and µ2, respectively, as in Fig. P1.48.
The plate spacings h1 and h2 are unequal, as shown. The
contact area is A between the center plate and each fluid.
(a) Assuming a linear velocity distribution in each fluid,
P1.49 An amazing number of commercial and laboratory devices have been developed to measure fluid viscosity, as
­described in Refs. 29 and 44. Consider a concentric
shaft, fixed axially and rotated inside the sleeve. Let the
inner and outer cylinders have radii ri and ro, respectively, with total sleeve length L. Let the rotational rate
be Ω (rad/s) and the applied torque be M. Using these
parameters, derive a theoretical relation for the viscosity
µ of the fluid between the cylinders.
P1.50 A simple viscometer measures the time t for a solid
sphere to fall a distance L through a test fluid of density
ρ. The fluid viscosity µ is then given by
μ≈
Wnett
3πDL
if
t≥
2ρDL
μ
where D is the sphere diameter and Wnet is the sphere net
weight in the fluid. (a) Prove that both of these formulas
are dimensionally homogeneous. (b) Suppose that a
2.5 mm diameter aluminum sphere (density 2700 kg/m3)
falls in an oil of density 875 kg/m3. If the time to fall 50 cm
is 32 s, estimate the oil viscosity and verify that the
­
inequality is valid.
P1.51 An approximation for the boundary-layer shape in
Figs. 1.6b and P1.51 is the formula
u( y) ≈ U sin (
πy
,
2δ )
0≤y≤δ
where U is the stream velocity far from the wall and δ is
the boundary-layer thickness, as in Fig. P1.51. If the fluid
is helium at 20°C and 1 atm, and if U = 10.8 m/s and
δ = 3 mm, use the formula to (a) estimate the wall shear
stress τw in Pa and (b) find the position in the boundary
layer where τ is one-half of τw.
52
Chapter 1 Introduction
*P1.54 A disk of radius R rotates at an angular velocity Ω inside
a disk-shaped container filled with oil of viscosity µ, as
shown in Fig. P1.54. Assuming a linear velocity profile
and neglecting shear stress on the outer disk edges,
­derive a formula for the viscous torque on the disk.
y
U
y=δ
Ω
Clearance
h
u(y)
0
P1.51
Oil
P1.52 The belt in Fig. P1.52 moves at a steady velocity V
and skims the top of a tank of oil of viscosity µ, as
shown. ­Assuming a linear velocity profile in the oil,
develop a simple formula for the required belt-drive
power P as a function of (h, L, V, b, µ). What beltdrive power P, in watts, is required if the belt moves
at 2.5 m/s over SAE 30W oil at 20°C, with L = 2 m,
b = 60 cm, and h = 3 cm?
V
L
Moving belt, width b
R
R
P1.54
P1.55 A block of weight W is being pulled over a table by another weight Wo, as shown in Fig. P1.55. Find an algebraic formula for the steady velocity U of the block if it
slides on an oil film of thickness h and viscosity µ. The
block bottom area A is in contact with the oil. Neglect
the cord weight and the pulley friction. Assume a linear
velocity profile in the oil film.
U
Oil, depth h
W
h
P1.52
*P1.53 A solid cone of angle 2θ, base r0, and density ρc is rotatWo
ing with initial angular velocity ω0 inside a conical seat, P1.55
as shown in Fig. P1.53. The clearance h is filled with oil
of viscosity µ. Neglecting air drag, derive an analytical *P1.56 The device in Fig. P1.56 is called a cone-plate viscometer [29]. The angle of the cone is very small, so that sin
­expression for the cone’s angular velocity ω(t) if there is
θ ≈ θ, and the gap is filled with the test liquid. The
no applied torque.
torque M to rotate the cone at a rate Ω is measured. Assuming a linear velocity profile in the fluid film, derive
ω (t)
Base
an expression for fluid viscosity µ as a function of (M,
radius r0
R, Ω, θ).
Oil
Ω
2θ
h
R
Fluid
θ
P1.53
P1.56
θ
Problems 53
P1.57 Extend the steady flow between a fixed lower plate and
a moving upper plate, from Fig. 1.8, to the case of two
­immiscible liquids between the plates, as in Fig. P1.57.
y
x
h2
µ2
h1
µ1
Fixed
P1.57
pig ­diameter is 5-15/16 in and its length 26 in. It cleans
a ­6-in-diameter pipe at a speed of 1.2 m/s. If the clearance is filled with glycerin at 20°C, what pressure difference, in pascals, is needed to drive the pig? Assume a
linear velocity profile in the oil and neglect air drag.
V
*P1.61 An air-hockey puck has a mass of 50 g and is 9 cm in
diameter. When placed on the air table, a 20°C air film,
of ­0.12-mm thickness, forms under the puck. The puck is
struck with an initial velocity of 10 m/s. Assuming a
linear velocity distribution in the air film, how long will
it take the puck to (a) slow down to 1 m/s and (b) stop
completely? Also, (c) how far along this extremely long
table will the puck have traveled for condition (a)?
(a) Sketch the expected no-slip velocity distribution u(y)
between the plates. (b) Find an analytic expression for
the velocity U at the interface between the two liquid
layers. (c) What is the result of (b) if the viscosities and
layer thicknesses are equal?
*P1.58 The laminar pipe flow example of Prob. 1.12 can be used
to design a capillary viscometer [29]. If Q is the volume
flow rate, L is the pipe length, and Δp is the pressure
drop from entrance to exit, the theory of Chap. 6 yields a
formula for viscosity:
μ=
πr40 Δp
8LQ
Pipe end effects are neglected [29]. Suppose our capillary
has r0 = 2 mm and L = 25 cm. The following flow rate
and pressure drop data are obtained for a certain fluid:
Q, m3/h
0.36
0.72
1.08
1.44
1.80
Δp, kPa
159
318
477
1274
1851
What is the viscosity of the fluid? Note: Only the first
three points give the proper viscosity. What is peculiar
about the last two points, which were measured accurately?
P1.59 A solid cylinder of diameter D, length L, and density ρs
falls due to gravity inside a tube of diameter D0. The
clearance, D0 − D << D, is filled with fluid of density ρ
and viscosity µ. Neglect the air above and below the cylinder. Derive a formula for the terminal fall velocity of
the cylinder. Apply your formula to the case of a steel
cylinder, D = 2 cm, D0 = 2.04 cm, L = 15 cm, with a
film of SAE 30 oil at 20°C.
P1.60 Pipelines are cleaned by pushing through them a close-­
fitting cylinder called a pig. The name comes from the
squealing noise it makes sliding along. Reference 50
­describes a new nontoxic pig, driven by compressed air,
for cleaning cosmetic and beverage pipes. Suppose the
Surface tension
P1.62 The hydrogen bubbles that produced the velocity profiles are quite small, D ≈ 0.01 mm. If the hydrogen–­
water interface is comparable to air–water and the water
temperature is 30°C, estimate the excess pressure within
the bubble.
P1.63 Derive Eq. (1.33) by making a force balance on the fluid
interface in Fig. 1.12c.
P1.64 Pressure in a water container can be measured by an
open vertical tube—see Fig. P2.11 for a sketch. If the
expected water rise is about 20 cm, what tube diameter is
needed to ensure that the error due to capillarity will be
less than 3 percent?
P1.65 The system in Fig. P1.65 is used to calculate the pressure
p1 in the tank by measuring the 15-cm height of liquid in
the 1-mm-diameter tube. The fluid is at 60°C. Calculate
the true fluid height in the tube and the percentage error
due to capillarity if the fluid is (a) water or (b) mercury.
15 cm
p1
P1.65
P1.66 DuNouy Tensiometer is a ring-pull device. It lifts a thin
wire ring from water. The force is very small and may be
measured by a calibrated soft-spring balance. Platinumiridium is recommended for the ring, being noncorrosive
and highly wetting to most liquids. A thin wire ring, 3 cm in
diameter, is lifted from a water surface at 20°C in the
54
Chapter 1 Introduction
d­ evice. Neglecting the wire weight, what is the force re- *P1.71 A soap bubble of diameter D1 coalesces with another bubble
quired to lift the ring?
of diameter D2 to form a single bubble D3 with the same
P1.67 A vertical concentric annulus, with outer radius ro and inamount of air. Assuming an isothermal process, derive an
expression for finding D3 as a function of D1, D2, patm, and Υ.
ner radius ri, is lowered into a fluid of surface tension Υ
and contact angle θ < 90°. Derive an expression for the
capillary rise h in the annular gap if the gap is very narrow.
Vapor pressure
*P1.68 Make an analysis of the shape η(x) of the water–air interface near a plane wall, as in Fig. P1.68, assuming that the P1.72 Early mountaineers boiled water to estimate their altitude. If they reach the top and find that water boils at
slope is small, R−1 ≈ d2η/dx2. Also assume that the pres84°C, ­approximately how high is the mountain?
sure difference across the interface is balanced by the
­specific weight and the interface height, Δp ≈ ρgη. The P1.73 A small submersible moves at velocity V, in fresh water
at 20°C, at a 2-m depth, where ambient pressure is
boundary conditions are a wetting contact angle θ at x =
131 kPa. Its critical cavitation number is known to be Ca
0 and a horizontal surface η = 0 as x → ∞. What is the
= 0.25. At what velocity will cavitation bubbles begin to
maximum height h at the wall?
form on the body? Will the body cavitate if V = 30 m/s
and the water is cold (5°C)?
y
P1.74 Oil, with a vapor pressure of 20 kPa, is delivered through
y=h
a pipeline by equally spaced pumps, each of which increases the oil pressure by 1.3 MPa. Friction losses in the
pipe are 150 Pa per meter of pipe. What is the maximum
possible pump spacing to avoid cavitation of the oil?
Speed of sound, Mach number
θ
η (x)
P1.68
x
x=0
P1.69 A solid cylindrical needle of diameter d, length L, and
density ρn may float in liquid of surface tension Y. Neglect buoyancy and assume a contact angle of 0°. Derive
a formula for the maximum diameter dmax able to float in
the liquid. Calculate dmax for a steel needle (SG = 7.84)
in water at 20°C.
P1.70 Derive an expression for the capillary height change h for
a fluid of surface tension Υ and contact angle θ between
two vertical parallel plates a distance W apart, as in Fig.
P1.70. What will h be for water at 20°C if W = 0.5 mm?
θ
h
P1.75 An airplane flies at 555 mi/h. At what altitude in the
standard atmosphere will the airplane’s Mach number be
­exactly 0.8?
P1.76 Derive a formula for the bulk modulus of an ideal gas, with
constant specific heats, and calculate it for steam at 300°C
and 200 kPa. Compare your result to the steam tables.
P1.77 Assume that the n-pentane data of Prob. P1.36 represent
isentropic conditions. Estimate the value of the speed of
sound at a pressure of 30 MPa. [Hint: The data approximately fit Eq. (1.19) with B = 260 and n = 11.]
P1.78 Sir Isaac Newton measured the speed of sound by timing
the difference between seeing a cannon’s puff of smoke
and hearing its boom. If the cannon is on a mountain 5.2
mi away, estimate the air temperature in degrees Celsius
if the time difference is (a) 24.2 s and (b) 25.1 s.
P1.79 From Table A.3, the density of glycerin at standard conditions is about 1260 kg/m3. At a very high pressure of
8000 lb/in2, its density increases to approximately 1275
kg/m3. Use these data to estimate the speed of sound of
glycerin, in ft/s.
P1.80 In Problem P1.24, for the given data, the air velocity at section 2 is 360 m/s. What is the Mach number at that ­section?
Streamlines
P1.81 Use Eq. (1.39) to find and sketch the streamlines of the
following flow field:
P1.70
W
u = Kx;
v = −Ky;
w = 0,
where K is a constant.
Comprehensive Problems 55
P1.82 A velocity field is given by u = V cosθ, v = V sinθ, and
w = 0, where V and θ are constants. Derive a formula for
the streamlines of this flow.
*P1.83 Use Eq. (1.39) to find and sketch the streamlines of the
­following flow field:
u = K(x2 − y2 ); v = −2Kxy; w = 0, where K is a constant.
Hint: This is a first-order exact differential equation.
History of fluid mechanics
P1.84 In the early 1900s, the British chemist Sir Cyril
­Hinshelwood quipped that fluid dynamics study was divided into “workers who observed things they could not
explain and workers who explained things they could not
observe.” To what historic situation was he referring?
P1.85 Do some reading and report to the class on the life and
achievements, especially vis-à-vis fluid mechanics, of
(a) Evangelista Torricelli (1608–1647)
(b) Henri de Pitot (1695–1771)
(c) Antoine Chézy (1718–1798)
(d) Gotthilf Heinrich Ludwig Hagen (1797–1884)
(e) Julius Weisbach (1806–1871)
(f) George Gabriel Stokes (1819–1903)
(g) Moritz Weber (1871–1951)
(h) Theodor von Kármán (1881–1963)
(i) Paul Richard Heinrich Blasius (1883–1970)
(j) Ludwig Prandtl (1875–1953)
(k) Osborne Reynolds (1842–1912)
(l) John William Strutt, Lord Rayleigh (1842–1919)
(m) Daniel Bernoulli (1700–1782)
(n) Leonhard Euler (1707–1783)
Fundamentals of Engineering Exam Problems
FE1.1 The absolute viscosity µ of a fluid is primarily a function of
(a) Density, (b) Temperature, (c) Pressure, (d) Velocity, (e) Surface tension
FE1.2 Carbon dioxide, at 20°C and 1 atm, is compressed isentropically to 4 atm. Assume CO2 is an ideal gas. The
­final temperature would be
(a) 130°C, (b) 162°C, (c) 171°C, (d) 237°C, (e) 313°C
FE1.3 Helium has a molecular weight of 4.003. What is the
weight of 2 m3 of helium at 1 atm and 20°C?
(a) 3.3 N, (b) 6.5 N, (c) 11.8 N, (d) 23.5 N, (e) 94.2 N
FE1.4 An oil has a kinematic viscosity of 1.25 E−4 m2/s and
a specific gravity of 0.80. What is its dynamic (absolute) viscosity in kg/(m · s)?
(a) 0.08, (b) 0.10, (c) 0.125, (d) 1.0, (e) 1.25
FE1.5 Consider a soap bubble of diameter 3 mm. If the surface
tension coefficient is 0.072 N/m and external pressure
is 0 Pa gage, what is the bubble’s internal gage pressure?
(a) −24 Pa, (b) +48 Pa, (c) +96 Pa, (d) +192 Pa,
(e) −192 Pa
FE1.6 The only possible dimensionless group that combines
velocity V, body size L, fluid density ρ, and surface
tension coefficient σ is
(a) Lρσ/V, (b) ρVL2/σ, (c) ρσV2/L, (d) σLV2/ρ,
(e) ρLV2/σ
Two parallel plates, one moving at 4 m/s and the other
fixed, are separated by a 5-mm-thick layer of oil of
specific gravity 0.80 and kinematic viscosity 1.25
E−4 m2/s. What is the average shear stress in the oil?
(a) 80 Pa, (b) 100 Pa, (c) 125 Pa, (d ) 160 Pa, (e) 200 Pa
FE1.8 Carbon dioxide has a specific-heat ratio of 1.30 and a
gas constant of 189 J/(kg · °C). If its temperature rises
from 20 to 45°C, what is its internal energy rise?
(a) 12.6 kJ/kg, (b) 15.8 kJ/kg, (c) 17.6 kJ/kg, (d) 20.5 kJ/
kg, (e) 25.1 kJ/kg
FE1.9 A certain water flow at 20°C has a critical cavitation
number, where bubbles form, Ca ≈ 0.25, where Ca =
2(pa− pvap)/ρV2. If pa = 1 atm and the vapor pressure
is 0.34 pounds per square inch absolute (psia), for
what ­water velocity will bubbles form?
(a) 12 mi/h, (b) 28 mi/h, (c) 36 mi/h, (d ) 55 mi/h,
(e) 63 mi/h
FE1.10 Example 1.10 gave an analysis that predicted that the
viscous moment on a rotating disk M = πµΩR4/(2h).
If the uncertainty of each of the four variables (µ, Ω,
R, h) is 1.0 percent, what is the estimated overall uncertainty of the moment M?
(a) 4.0 percent (b) 4.4 percent (c) 5.0 percent
(d ) 6.0 percent (e) 7.0 percent
FE1.7
Comprehensive Problems
C1.1
Sometimes we can develop equations and solve practical problems by knowing nothing more than the dimensions of the key parameters in the problem. For example,
consider the heat loss through a window in a building.
Window efficiency is rated in terms of “R value,” which
has units of (ft2 · h · °F)/Btu. A certain manufacturer advertises a double-pane window with an R value of 2.5.
The same company produces a triple-pane window with
an R value of 3.4. In either case the window dimensions
are 3 ft by 5 ft. On a given winter day, the temperature
56
C1.2
C1.3
Chapter 1 Introduction
difference between the inside and outside of the building is 45°F.
(a) Develop an equation for the amount of heat lost in a
given time period Δt, through a window of area A,
with a given R value, and temperature difference
ΔT. How much heat (in Btu) is lost through the double-pane window in one 24-h period?
(b) How much heat (in Btu) is lost through the triplepane window in one 24-h period?
(c) Suppose the building is heated with propane gas,
which costs $3.25 per gallon. The propane burner is
80 percent efficient. Propane has approximately
90,000 Btu of available energy per gallon. In that
same 24-h ­period, how much money would a homeowner save per window by installing triple-pane
rather than double-pane windows?
(d) Finally, suppose the homeowner buys 20 such triplepane windows for the house. A typical winter has the
equivalent of about 120 heating days at a temperature
difference of 45°F. Each triple-pane window costs
$85 more than the double-pane window. Ignoring interest and inflation, how many years will it take the
homeowner to make up the additional cost of the triple-pane windows from heating bill savings?
When a person ice skates, the surface of the ice actually
melts beneath the blades, so that he or she skates on a
thin sheet of water between the blade and the ice.
(a) Find an expression for total friction force on the
­bottom of the blade as a function of skater velocity V,
blade length L, water thickness (between the blade
and the ice) h, water viscosity µ, and blade width W.
(b) Suppose an ice skater of total mass m is skating along
at a constant speed of V0 when she suddenly stands
stiff with her skates pointed directly forward, allowing herself to coast to a stop. Neglecting friction due
to air resistance, how far will she travel before she
comes to a stop? (Remember, she is coasting on two
skate blades.) Give your answer for the total distance
traveled, x, as a function of V0, m, L, h, µ, and W.
(c) Find x for the case where V0 = 4.0 m/s, m = 100 kg, L =
30 cm, W = 5.0 mm, and h = 0.10 mm. Do you think our
assumption of negligible air resistance is a good one?
Two thin flat plates, tilted at an angle α, are placed in a
tank of liquid of known surface tension Υ and contact
angle θ, as shown in Fig. C1.3. At the free surface of the
liquid in the tank, the two plates are a distance L apart
and have width b into the page. The liquid rises a distance h between the plates, as shown.
(a) What is the total upward (z-directed) force, due to
surface tension, acting on the liquid column between
the plates?
(b) If the liquid density is ρ, find an expression for surface tension Υ in terms of the other variables.
α
α
θ
θ
h
z
g
L
C1.3
C1.4
Oil of viscosity µ and density ρ drains steadily down the
side of a tall, wide vertical plate, as shown in Fig. C1.4.
In the region shown, fully developed conditions exist;
that is, the velocity profile shape and the film thickness
δ are independent of distance z along the plate. The vertical velocity w becomes a function only of x, and the
shear resistance from the atmosphere is negligible.
Plate
Oil film
Air
δ
g
z
C1.4
x
(a) Sketch the approximate shape of the velocity profile
w(x), considering the boundary conditions at the
wall and at the film surface.
(b) Suppose film thickness, δ, and the slope of the
­velocity profile at the wall, (dw/dx)wall, are measured
by a laser Doppler anemometer (to be discussed in
Chap. 6). Find an expression for the viscosity of the
oil as a function of ρ, δ, (dw/dx)wall, and the gravitational acceleration g. Note that, for the coordinate
system given, both w and (dw/dx)wall are negative.
Comprehensive Problems 57
C1.5
C1.6
Viscosity can be measured by flow through a thin-bore
or capillary tube if the flow rate is low. For length L,
(small) diameter D ≪ L, pressure drop Δp, and (low)
volume flow rate Q, the formula for viscosity is µ =
D4Δp/(CLQ), where C is a constant.
(a) Verify that C is dimensionless. The following data
are for water flowing through a 2-mm-diameter tube
which is 1 meter long. The pressure drop is held
constant at Δp = 5 kPa.
T, °C
10.0
40.0
70.0
Q, L/min
0.091
0.179
0.292
(b) Using proper SI units, determine an average value of
C by accounting for the variation with temperature
of the viscosity of water.
The rotating-cylinder viscometer in Fig. C1.6 shears the
fluid in a narrow clearance Δr, as shown. Assuming a
linear velocity distribution in the gaps, if the driving
torque M is measured, find an expression for µ by (a)
neglecting and (b) including the bottom friction.
Ω
R
L
where A and B are constants that are determined by calibrating the device with a known fluid. Here are calibration data for a Stormer viscometer tested in glycerol,
using a weight of 50 N:
C1.9
µ, kg/(m-s)
0.23
0.34
0.57
0.84
1.15
t, sec
15
23
38
56
77
(a) Find reasonable values of A and B to fit these calibration data. Hint: The data are not very sensitive to
the value of B.
(b) A more viscous fluid is tested with a 100 N weight
and the measured time is 44 s. Estimate the viscosity
of this fluid.
The lever in Fig. C1.9 has a weight W at one end and is
tied to a cylinder at the left end. The cylinder has negligible weight and buoyancy and slides upward through a
film of heavy oil of viscosity µ. (a) If there is no acceleration (uniform lever rotation), derive a formula for the
rate of fall V2 of the weight. Neglect the lever weight.
Assume a linear velocity profile in the oil film. (b) Estimate the fall velocity of the weight if W = 20 N, L1 = 75
cm, L2 = 50 cm, D = 10 cm, L = 22 cm, ΔR = 1 mm, and
the oil is glycerin at 20°C.
Viscous
fluid µ
L1
V1
Solid
cylinder
∆r << R
C1.8
Make an analytical study of the transient behavior of the
sliding block in Prob. 1.45. (a) Solve for V(t) if the block
starts from rest, V = 0 at t = 0. (b) Calculate the time t1
when the block has reached 98 percent of its terminal
­velocity.
A mechanical device that uses the rotating cylinder of
Fig. C1.6 is the Stormer viscometer [29]. Instead of being driven at constant Ω, a cord is wrapped around the
shaft and attached to a falling weight W. The time t to
turn the shaft a given number of revolutions (usually
five) is measured and correlated with viscosity. The formula is
Aμ
t≈
W−B
pivot
W
V2?
Cylinder, diameter D, length L,
in an oil film of thickness ΔR.
C1.6
C1.7
L2
C1.9
C1.10 A popular gravity-driven instrument is the Cannon­
Ubbelohde viscometer, shown in Fig. C1.10. The test
­liquid is drawn up above the bulb on the right side and
­allowed to drain by gravity through the capillary tube
­below the bulb. The time t for the meniscus to pass from
upper to lower timing marks is recorded. The kinematic
viscosity is computed by the simple formula:
v = Ct
where C is a calibration constant. For ν in the range of
100–500 mm2/s, the recommended constant is C =
0.50 mm2/s2, with an accuracy less than 0.5 percent.
58
Chapter 1 Introduction
ball of diameter D and density ρb falls through a tube of
test liquid (ρ, µ). The fall velocity V is calculated by the
time to fall a measured distance. The formula for calculating the viscosity of the fluid is
Upper timing mark
Lower timing mark
Capillary tube
C1.10 The CannonUbbelohde viscometer.
Cannon Instrument
Company.
(a) What liquids from Table A.3 are in this viscosity
range? (b) Is the calibration formula dimensionally consistent? (c) What system properties
might the constant C depend upon? (d ) What
problem in this chapter hints at a formula for estimating the viscosity?
C1.11 Mott [Ref. 44, p. 38] discusses a simple falling-ball viscometer, which we can analyze later in Chap. 7. A small
μ=
(ρb − ρ)gD2
18 V
This result is limited by the requirement that the Reynolds number (ρVD/µ) be less than 1.0. Suppose a steel ball
(SG = 7.87) of diameter 2.2 mm falls in SAE 25W oil
(SG = 0.88) at 20°C. The measured fall velocity is
8.4 cm/s. (a) What is the viscosity of the oil, in kg/m-s?
(b) Is the Reynolds number small enough for a valid estimate?
C1.12 A solid aluminum disk (SG = 2.7) is 2 in in diameter and
3/16 in thick. It slides steadily down a 14° incline that is
coated with a castor oil (SG = 0.96) film one hundredth
of an inch thick. The steady slide velocity is 2 cm/s. Using Figure A.1 and a linear oil velocity profile assumption, ­estimate the temperature of the castor oil.
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J. Hilsenrath et al., “Tables of Thermodynamic and Transport Properties,” U.S. Nat. Bur. Standards Circular 564,
1955; reprinted by Pergamon, New York, 1960.
W. T. Parry, ASME International Steam Tables for Industrial Use, 2d ed., ASME Press, New York, 2009.
Steam Tables URL: http://www.steamtablesonline.com/
O. A. Hougen and K. M. Watson, Chemical Process Principles Charts, Wiley, New York, 1960.
F. M. White, Viscous Fluid Flow, 3d ed., McGraw-Hill,
New York, 2005.
M. Bourne, Food Texture and Viscosity: Concept and Measurement, 2d ed., Academic Press, Salt Lake City, Utah, 2002.
SAE Fuels and Lubricants Standards Manual, Society of
­Automotive Engineers, Warrendale, PA, 2001.
C. L. Yaws, Handbook of Viscosity, 3 vols., Elsevier
­Science and Technology, New York, 1994.
A. W. Adamson and A. P. Gast, Physical Chemistry of
­Surfaces, Wiley, New York, 1999.
C. E. Brennen, Fundamentals of Multiphase Flow,
­Cambridge University Press, New York, 2009.
National Committee for Fluid Mechanics Films, Illustrated
Experiments in Fluid Mechanics, M.I.T. Press, Cambridge,
MA, 1972.
I. G. Currie, Fundamental Mechanics of Fluids, 3d ed.,
­Marcel Dekker, New York, 2003.
W.-J. Yang (ed.), Handbook of Flow Visualization, 2d ed.,
Taylor and Francis, New York, 2001.
F. T. Nieuwstadt (ed.), Flow Visualization and Image
­Analysis, Springer, New York, 2007.
A. J. Smits and T. T. Lim, Flow Visualization: Techniques
and Examples, 2d ed., Imperial College Press, London, 2011.
R. J. Adrian and J. Westerweel, Particle Image Velocimetry, Cambridge University Press, New York, 2010.
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39.
40.
41.
42.
43.
44.
45.
46.
47.
48.
49.
Wen-Jai Yang, Computer-Assisted Flow Visualization,
­Begell House, New York, 1994.
M. van Dyke, An Album of Fluid Motion, Parabolic Press,
Stanford, CA, 1982.
Y. Nakayama and Y. Tanida (eds.), Visualized Flow, vol. 1,
Elsevier, New York, 1993; vols. 2 and 3, CRC Press, Boca
Raton, FL, 1996.
M. Samimy, K. S. Breuer, L. G. Leal, and P. H. Steen, A
Gallery of Fluid Motion, Cambridge University Press, New
York, 2003.
NASA Langley Research Center, “Wake Vortex Study at
Wallops Island,” URL https://en.wikipedia.org/wiki/Wingtip_vortices#/media/File:Airplane_vortex_edit.jpg.
B. Carr and V. E. Young, “Videotapes and Movies on
Fluid Dynamics and Fluid Machines,” in Handbook of
Fluid ­Dynamics and Fluid Machinery, vol. II, J. A. Schetz
and A. E. Fuhs (eds.), Wiley, New York, 1996, pp. 1171–
1189.
R. L. Mott, Applied Fluid Mechanics, Pearson PrenticeHall, Upper Saddle River, NJ, 2006.
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Plane,” J. Fluid Mechanics, vol. 20, 1964, pp. 481–499.
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­Viscosity of n-alkanes,” Int. J. Thermophysics, vol. 13, no.
3, 1992, pp. 411–442.
60
The bathyscaphe Trieste was built by explorers August and Jacques Picard in 1953. They made
many deep-sea dives over the next five years. In 1958 the U.S. Navy purchased it and, in 1960,
descended to the deepest part of the ocean, the Marianas Trench, near Guam. The recorded
depth was 10,900 m, or nearly seven miles. The photo shows the vessel being lifted from the
water. The pressure sphere is on the bottom, and the boat-shaped structure above is filled with
gasoline, which provides buoyancy even at those depths. [Image courtesy of the U.S. Navy.]
Chapter 2
Pressure Distribution
in a Fluid
Motivation. Many fluid problems do not involve motion. They concern the pres-
sure distribution in a static fluid and its effect on solid surfaces and on floating
and submerged bodies.
When the fluid velocity is zero, denoted as the hydrostatic condition, the pressure variation is due only to the weight of the fluid. Assuming a known fluid in a
given gravity field, the pressure may easily be calculated by integration. Important
applications in this chapter are (1) pressure distribution in the atmosphere and the
oceans, (2) the design of manometer, mechanical, and electronic pressure instruments, (3) forces on submerged flat and curved surfaces, (4) buoyancy on a submerged body, and (5) the behavior of floating bodies. The last two result in
Archimedes’ principles.
If the fluid is moving in rigid-body motion, such as a tank of liquid that has
been spinning for a long time, the pressure also can be easily calculated because
the fluid is free of shear stress. We apply this idea here to simple rigid-body
accelerations in Sec. 2.9. Pressure measurement instruments are discussed in Sec.
2.10. As a matter of fact, pressure also can be analyzed in arbitrary (nonrigid-body)
motions V(x, y, z, t), but we defer that subject to Chap. 4.
2.1 Pressure and Pressure Gradient
In Fig. 1.1 we saw that a fluid at rest cannot support shear stress, and thus Mohr’s
circle reduces to a point. In other words, the normal stress on any plane through
a fluid element at rest is a point property called the fluid pressure p, taken
positive for compression by common convention. This is such an important concept that we shall review it with another approach.
First let us emphasize that pressure is a thermodynamic property of the fluid,
like temperature or density. It is not a force. Pressure has no direction and is not
a vector. The concept of force only arises when considering a surface immersed
in fluid ­pressure. The pressure creates a force, due to fluid molecules bombarding
the surface, and it is normal to that surface.
61
62
Chapter 2 Pressure Distribution in a Fluid
z (up)
pn
θ
∆s
Element weight:
d W = ρg ( 12 b ∆x ∆z)
∆z
px
∆x
θ
x
O
Fig. 2.1 Equilibrium of a small
wedge of fluid at rest.
Width b into paper
pz
Figure 2.1 shows a small wedge of fluid at rest of size Δx by Δz by Δs and
depth b into the paper. There is no shear by definition, but we postulate that
the pressures px, pz, and pn may be different on each face. The weight of the
element also may be important. The element is assumed to be small, so the
pressure is constant on each face. Summation of forces must equal zero (no
acceleration) in both the x and z directions.
gFx = 0 = px b Δz − pnb Δs sin θ
gFz = 0 = pzb Δx − pnb Δs cos θ − 12ρgb Δx Δz(2.1)
But the geometry of the wedge is such that
Δs sin θ = Δz
Δs cos θ = Δx(2.2)
Substitution into Eq. (2.1) and rearrangement give
px = pn
pz = pn + 12ρg Δz(2.3)
These relations illustrate two important principles of the hydrostatic, or shear-free,
condition: (1) There is no pressure change in the horizontal direction, and (2)
there is a vertical change in pressure proportional to the density, gravity, and
depth change. We shall exploit these results to the fullest, starting in Sec. 2.3.
In the limit as the fluid wedge shrinks to a “point,” Δz → 0 and Eq. (2.3)
become
px = pz = pn = p(2.4)
Since θ is arbitrary, we conclude that the pressure p in a static fluid is a point
property, independent of orientation.
Pressure Force on a Fluid Element
Pressure (or any other stress, for that matter) causes a net force on a fluid element. To see this, consider the pressure acting on the two x faces in Fig. 2.2. Let
the ­pressure vary arbitrarily
p = p(x, y, z, t)
2.2 Equilibrium of a Fluid Element 63
z
dx
pdxdz
(p +
dz
∂p
d y) dx dz
∂y
y
Fig. 2.2 Net y force on an element
due to pressure variation.
x
dy
The net force in the y direction on the element in Fig. 2.2 is given by
∂p
∂p
dFy = p dx dz − (p +
dy) dx dz = − dx dy dz
∂y
∂y
In like manner, the net force dFx involves −∂p/∂x, and the net force dFz concerns
−∂p/∂z. The total net-force vector on the element due to pressure is
dFpress = (−i
∂p
∂p
∂p
−j
− k ) dx dy dz(2.5)
∂x
∂y
∂z
We recognize the term in parentheses as the negative vector gradient of p. Denoting f as the net force per unit element volume, we rewrite Eq. (2.5) as
fpress = −∇p(2.6)
where
∇ = gradient operator = i
∂
∂
∂
+j
+k
∂x
∂y
∂z
Thus it is not the pressure but the pressure gradient causing a net force that must
be balanced by gravity or acceleration or some other effect in the fluid.
2.2 Equilibrium of a Fluid Element
The pressure gradient is a surface force that acts on the sides of the element.
There may also be a body force, due to electromagnetic or gravitational potentials,
acting on the entire mass of the element. Here we consider only the gravity force,
or weight of the element:
d Fgrav = ρg dx dy dz
or
fgrav = ρg
(2.7)
In addition to gravity, a fluid in motion will have surface forces due to viscous
stresses, which are also called shear stresses. By Newton’s law, Eq. (1.2), the sum
of these per-unit-volume forces equals the mass per unit volume (density) times
the acceleration a of the fluid element:
∑ f = fpress + fgrav + fvisc = −∇p + ρg + fvisc = ρa(2.8)
64
Chapter 2 Pressure Distribution in a Fluid
p (Pascals)
120,000
30,000
90,000
High pressure:
p = 120,000 Pa abs = 30,000 Pa gage
Local atmosphere:
p = 90,000 Pa abs = 0 Pa gage = 0 Pa vacuum
40,000
Vacuum pressure:
p = 50,000 Pa abs = 40,000 Pa vacuum
50,000
50,000
Fig. 2.3 Illustration of absolute,
gage, and vacuum pressure readings.
0
Absolute zero reference:
p = 0 Pa abs = 90,000 Pa vacuum
(Tension)
This general equation will be studied in detail in Chap. 4. Note that Eq. (2.8) is
a vector relation, and the acceleration may not be in the same vector direction as
the velocity. For our present topic, hydrostatics, the viscous stresses and the
acceleration are zero.
Gage Pressure and Vacuum Pressure: Relative Terms
Before embarking on examples, we should note that engineers are apt to specify
pressures as (1) the absolute or total magnitude or (2) the value relative to the local
ambient atmosphere. The second case occurs because many pressure instruments
are of differential type and record, not an absolute magnitude, but the difference
between the fluid pressure and the atmosphere. The measured pressure may be
either higher or lower than the local atmosphere, and each case is given a name:
1. p > pa Gage pressure:
2. p < pa Vacuum pressure:
p(gage) = p – pa
p(vacuum) = pa – p
This is a convenient shorthand, and one later adds (or subtracts) atmospheric
pressure to determine the absolute fluid pressure.
A typical situation is shown in Fig. 2.3. The local atmosphere is at, say, 90,000 Pa,
which might reflect a storm condition in a sea-level location or normal conditions
at an altitude of 1000 m. Thus, on this day, pa = 90,000 Pa absolute = 0 Pa
gage = 0 Pa vacuum. Suppose gage 1 in a laboratory reads p1 = 120,000 Pa absolute. This value may be reported as a gage pressure, p1 = 120,000 − 90,000 =
30,000 Pa gage. (One must also record the atmospheric pressure in the laboratory,
since pa changes gradually.) Suppose gage 2 reads p2 = 50,000 Pa absolute. Locally,
this is a vacuum pressure and might be reported as p2 = 90,000 − 50,000 = 40,000
Pa vacuum. Occasionally, in the problems section, we will specify gage or vacuum
pressure to keep you alert to this common engineering practice. If a pressure is
listed without the modifier gage or vacuum, we assume it is absolute pressure.
2.3 Hydrostatic Pressure Distributions
If the fluid is at rest or at constant velocity, a = 0 and fvisc = 0. Equation (2.8)
for the pressure distribution reduces to
∇p = ρg(2.9)
2.3 Hydrostatic Pressure Distributions 65
This is a hydrostatic distribution and is correct for all fluids at rest, regardless of
their viscosity, because the viscous term vanishes identically.
Recall from vector analysis that the vector ∇p expresses the magnitude and
direction of the maximum spatial rate of increase of the scalar property p. As a
result, ∇p is perpendicular everywhere to surfaces of constant p. Thus Eq. (2.9)
states that a fluid in hydrostatic equilibrium will align its constant-pressure surfaces everywhere normal to the local-gravity vector. The maximum pressure
increase will be in the direction of gravity—that is, “down” in the negative z
direction. If the fluid is a liquid, its free surface, being at atmospheric pressure,
will be normal to local gravity, or “horizontal.” You probably knew all this before,
but Eq. (2.9) is the proof of it.
The local-gravity vector can be expressed as
g = −gk(2.10)
where g is the magnitude of local gravity, for example, 9.807 m/s2. For these
coordinates Eq. (2.9) has the components
∂p
=0
∂x
∂p
=0
∂y
∂p
= −ρg = −γ(2.11)
∂z
the first two of which tell us that p is independent of x and y. Hence ∂p/∂z can
be replaced by the total derivative dp/dz, and the hydrostatic condition reduces to
dp
= −γ
dz
or
p2 − p1 = −
2
∫ γ dz(2.12)
1
Equation (2.12) is the solution to the hydrostatic problem. The integration requires
an assumption about the density and gravity distribution. Gases and liquids are
usually treated differently.
We state the following conclusions about a hydrostatic condition:
Pressure in a continuously distributed uniform static fluid varies only with
vertical distance and is independent of the shape of the container. The pressure
is the same at all points on a given horizontal plane in the fluid. The pressure
increases with depth in the fluid.
An illustration of this is shown in Fig. 2.4. The free surface of the container is
atmospheric and forms a horizontal plane. Points a, b, c, and d are at equal depth
in a horizontal plane and are interconnected by the same fluid, water; therefore,
all points have the same pressure. The same is true of points A, B, and C on the
bottom, which all have the same higher pressure than at a, b, c, and d. However,
point D, although at the same depth as A, B, and C, has a different pressure
because it lies beneath a different fluid, mercury.
Effect of Variable Gravity
For a spherical planet of uniform density, the acceleration of gravity varies
inversely as the square of the radius from its center
r0 2
g = g0 ( )
r
(2.13)
66
Chapter 2 Pressure Distribution in a Fluid
Atmospheric pressure
Fig. 2.4 Hydrostatic-pressure distribution. Points a, b, c, and d are at
equal depths in water and therefore
have identical pressures. Points A,
B, and C are also at equal depths in
water and have identical pressures
higher than a, b, c, and d. Point D
has a different pressure from A, B,
and C because it is not connected to
them by a water path.
Free surface
Water
Depth 1
a
b
c
d
Mercury
Depth 2
A
B
C
D
where r0 is the planet radius and g0 is the surface value of g. For earth, r0 ≈ 3960
statute mi ≈ 6400 km. In typical engineering problems, the deviation from r0 extends
from the deepest ocean, about 11 km, to the atmospheric height of supersonic transport operation, about 20 km. This gives a maximum variation in g of (6400/6420)2,
or 0.6 percent. We therefore neglect the variation of g in most problems.
Hydrostatic Pressure in Liquids
Liquids are so nearly incompressible that we can neglect their density variation
in hydrostatics. In Example 1.6 we saw that water density increases only 4.6
percent at the deepest part of the ocean. Its effect on hydrostatics would be about
half of this, or 2.3 percent. Thus we assume constant density in liquid hydrostatic
calculations, for which Eq. (2.12) integrates to
Liquids:
p2 − p1 = −γ (z2 − z1 )
or
z1 − z2 =
(2.14)
p2 p1
−
γ
γ
We use the first form in most problems. The quantity γ is called the specific weight
of the fluid, with dimensions of weight per unit volume; some values are tabulated
in Table 2.1. The quantity p/γ is a length called the pressure head of the fluid.
Table 2.1 Specific Weight of
Some Common Fluids
Specific weight γ
at 68°F = 20°C
Fluid
Air (at 1 atm)
Ethyl alcohol
SAE 30 oil
Water
Seawater
Glycerin
Carbon tetrachloride
Mercury
lbf/ft3
N/m3
0.0752
49.2
55.5
62.4
64.0
78.7
99.1
846
11.8
7,733
8,720
9,790
10,050
12,360
15,570
133,100
2.3 Hydrostatic Pressure Distributions 67
Z
+b
p ≈ pa – bγair
Air
Free surface: Z = 0, p = pa
0
Fig. 2.5 Hydrostatic-pressure
distribution in oceans and
atmospheres.
Water
g
–h
p ≈ pa + hγwater
For lakes and oceans, the coordinate system is usually chosen as in Fig. 2.5,
with z = 0 at the free surface, where p equals the surface atmospheric pressure
pa. When we introduce the reference value (p1, z1) = (pa, 0), Eq. (2.14) becomes,
for p at any (negative) depth z,
Lakes and oceans:
p = pa − γz(2.15)
where γ is the average specific weight of the lake or ocean. As we shall see,
Eq. (2.15) holds in the atmosphere also with an accuracy of 2 percent for heights
z up to 1000 m.
EXAMPLE 2.1
Newfound Lake, a freshwater lake near Bristol, New Hampshire, has a maximum depth
of 60 m, and the mean atmospheric pressure is 91 kPa. Estimate the absolute pressure
in kPa at this maximum depth.
Solution
∙ System sketch: Imagine that Fig. 2.5 is Newfound Lake, with h = 60 m and z = 0
at the surface.
∙ Property values: From Table 2.1, γwater = 9790 N/m3. We are given that patmos = 91 kPa.
∙ Solution steps: Apply Eq. (2.15) to the deepest point. Use SI units, pascals, not kilopascals:
pmax = pa − γ z = 91,000 Pa − (9790
N
(−60 m) = 678,400 Pa ≈ 678 kPa Ans.
m3 )
∙ Comments: Kilopascals are awkward. Use pascals in the formula, then convert the answer.
The Mercury Barometer
The simplest practical application of the hydrostatic formula (2.14) is the barometer
(Fig. 2.6), which measures atmospheric pressure. A tube is filled with mercury
and inverted while submerged in a reservoir. This causes a near vacuum in the closed
upper end because mercury has an extremely small vapor pressure at room temperatures
68
Chapter 2 Pressure Distribution in a Fluid
p1 ≈ 0
(Mercury has a very
low vapor pressure.)
z1 = h
p2 ≈ pa
(The mercury is in
contact with the
atmosphere.)
p
h= γa
M
z
pa
z2 = 0
pM
Mercury
(a)
(b)
Fig. 2.6 A barometer measures local absolute atmospheric pressure: (a) the height of a
mercury column is proportional to patm; (b) A handheld digital barometer-altimeter. (Courtesy
of Nova Lynx Corporation)
(0.16 Pa at 20˚C). Since atmospheric pressure forces a mercury column to rise a
distance h into the tube, the upper mercury surface is at zero pressure.
From Fig. 2.6, Eq. (2.14) applies with p1 = 0 at z1 = h and p2 = pa at z2 = 0:
or
pa − 0 = −γM (0 − h)
pa
h = (2.16)
γM
At sea-level standard, with pa = 101,350 Pa and γM = 133,100 N/m3 from Table
2.1, the barometric height is h = 101,350/133,100 = 0.761 m or 761 mm. In the
United States the weather service reports this as an atmospheric “pressure” of
29.96 inHg (inches of mercury). Mercury is used because it is the heaviest common liquid. A water barometer would be 34 ft high.
Hydrostatic Pressure in Gases
Gases are compressible, with density nearly proportional to pressure. Thus density
must be considered as a variable in Eq. (2.12) if the integration carries over large
pressure changes. It is sufficiently accurate to introduce the perfect-gas law p =
ρRT in Eq. (2.12):
dp
p
= −ρg = −
g
dz
RT
2.3 Hydrostatic Pressure Distributions 69
Separate the variables and integrate between points 1 and 2:
∫
2
1
p2
dp
g
= ln = −
p
p1
R
∫
2
1
dz
(2.17)
T
The integral over z requires an assumption about the temperature variation T(z).
One common approximation is the isothermal atmosphere, where T = T0:
p2 = p1 exp [ −
g(z2 − z1 )
(2.18)
RT0 ]
The quantity in brackets is dimensionless. (Think that over; it must be dimensionless, right?) Equation (2.18) is a fair approximation for earth, but actually the
earth’s mean atmospheric temperature drops off nearly linearly with z up to an
altitude of about 36,000 ft (11,000 m):
T ≈ T0 − Bz
(2.19)
Here T0 is sea-level temperature (absolute) and B is the lapse rate, both of which
vary somewhat from day to day.
The Standard Atmosphere
By international agreement [1], the following standard values are assumed to
apply from 0 to 36,000 ft:
T0 = 518.69°R = 288.16 K = 15°C
B = 0.003566°R/ft = 0.00650 K /m
This lower portion of the atmosphere is called the troposphere. Introducing
Eq. (2.19) into Eq. (2.17) and integrating, we obtain the more accurate relation
p = pa (1 −
Bz g/(RB)
T0 )
where
g
= 5.26 (air)
RB
g
Bz RB −1
ρ = ρo(1 − )
To
where ρo = 1.2255
kg
m3
(2.20)
, po = 101,350 pa in the troposphere, with z = 0 at sea level. The exponent g/(RB) is dimensionless (again
it must be) and has the standard value of 5.26 for air, with R = 287 m2/(s2 · K).
The U.S. standard atmosphere [1] is sketched in Fig. 2.7. The pressure is seen
to be nearly zero at z = 30 km. For tabulated properties, see Table A.6.
EXAMPLE 2.2
If sea-level pressure is 101,350 Pa, compute the standard pressure at an altitude of 5000
m, using (a) the exact formula and (b) an isothermal assumption at a standard sea-level
temperature of 15˚C. Is the isothermal approximation adequate?
70
Chapter 2 Pressure Distribution in a Fluid
Solution
Part (a)
Use absolute temperature in the exact formula, Eq. (2.20):
p = pa [ 1 −
(0.00650 K/m) (5000 m) 5.26
5.26
] = (101,350 Pa) (0.8872)
288.16 K
= 101,350(0.5328) = 54,000 Pa
Ans. (a)
This is the standard-pressure result given at z = 5000 m in Table A.6.
Part (b)
If the atmosphere were isothermal at 288.16 K, Eq. (2.18) would apply:
p ≈ pa exp (−
gz
(9.807 m/s2 ) (5000 m)
= (101,350 Pa) exp {−
)
RT
[287 m2/(s2 · K) ] (288.16 K)}
= (101,350 Pa) exp(−0.5929) ≈ 56,000 Pa
Ans. (b)
This is 4 percent higher than the exact result. The isothermal formula is inaccurate in
the troposphere.
Is the Linear Formula Adequate for Gases?
The linear approximation from Eq. (2.14), δp ≈ –ρg δz, is satisfactory for liquids,
which are nearly incompressible. For gases, it is inaccurate unless δz is rather
small. Problem P2.26 asks you to show, by binomial expansion of Eq. (2.20), that
the error in using constant gas density to estimate δp from Eq. (2.14) is small if
Source: U.S. Standard Atmosphere,
1976, Government Printing Office,
Washington DC, 1976.
(2.21)
50
50
40
40
Altitude z, km
60
30
20
Fig. 2.7 Temperature and pressure
distribution in the U.S. standard
atmosphere.
2T0
(n − 1)B
60
–56.5°C
Altitude z, km
δz ≪
10
– 60
30
20.1 km
20
11.0 km
10
– 40
Eq. (2.18)
Eq. (2.20)
Eq. (2.19)
Troposphere
0
1.20 kPa
101.33 kPa
15°C
– 20
Temperature, °C
0
+20
0
40
80
Pressure, kPa
120
2.4 Application to Manometry 71
where T0 is the local absolute temperature, B is the lapse rate from Eq. (2.19),
and n = g/(RB) is the exponent in Eq. (2.20). The error is less than 1 percent if
δz < 200 m.
2.4 Application to Manometry
From the hydrostatic formula (2.14), a change in elevation z2 – z1 of a liquid
is equivalent to a change in pressure (p2 – p1)/γ. Thus a static column of one
or more liquids or gases can be used to measure pressure differences between
two points. Such a device is called a manometer. Because the manometer
typically consists of a tube formed into the shape of a U, it is often called
the U-tube manometer. If multiple fluids are used, we must change the density
in the formula as we move from one fluid to another. Figure 2.8 illustrates
the use of the formula with a column of multiple fluids. The pressure change
through each fluid is calculated separately. If we wish to know the total
change p5 – p1, we add the successive changes p2 – p1, p3 – p2, p4 – p3, and
p5 – p4. The intermediate values of p cancel, and we have, for the example
of Fig. 2.8,
p5 − p1 = −γ0 (z2 − z1 ) − γw (z3 − z2 ) − γG (z4 − z3 ) − γM (z5 − z4 )(2.22)
No additional simplification is possible on the right-hand side because of the
different densities. Notice that we have placed the fluids in order from the lightest
on top to the heaviest at bottom. This is the only stable configuration. If we
attempt to layer them in any other manner, the fluids will overturn and seek the
stable arrangement.
Pressure Increases Downward
The basic hydrostatic relation, Eq. (2.14), is mathematically correct but vexing to
engineers because it combines two negative signs to have the pressure increase
downward. When calculating hydrostatic pressure changes, engineers work
instinctively by simply having the pressure increase downward and decrease
upward. For the multiple fluid layers in Fig. 2.8, we can start at point z5 and work
z = z1
z2
z
z3
z4
Fig. 2.8 Evaluating pressure
changes through a column of
multiple fluids.
z5
Known pressure p1
Oil, ρo
Water, ρw
Glycerin, ρG
Mercury, ρM
p2 – p1 = – ρo g(z 2 – z1)
p3 – p2 = – ρw g(z 3 – z 2)
p4 – p3 = – ρG g(z 4 – z 3)
p5 – p4 = – ρ g(z – z )
M
5
4
Sum = p5 – p1
72
Chapter 2 Pressure Distribution in a Fluid
our way to the top layer at z1 with known pressure. As we move from point z5 to
z4, the pressure will decrease by γM(z4 – z5). Similarly, we will move up to z3, z2,
and finally z1. In equation form, these steps can be expressed as
p5 − γM(z4 − z5) − γG(z3 − z4) − γw(z2 − z3) − γo(z1 − z2) = p1
That is, keep subtracting on pressure decrements as you move up through the
layered fluid. This method is particularly useful when it is applied to a manometer,
which involves both “up” and “down” calculations.
Application: A Simple Manometer
Figure 2.9 shows a simple U-tube open manometer that measures the gage pressure pA relative to the atmosphere, pa. The chamber fluid ρ1 is separated from the
atmosphere by a second, heavier fluid ρ2, perhaps because fluid A is corrosive,
or more likely because a heavier fluid ρ2 will keep z2 small and the open tube
can be shorter.
We first start at A and work around to the open end. The pressure at A and
point (1) are the same, and when we move from point (1) to (2), the pressure
increases by γ1∣zA – z1∣. The pressure at point (2) is equal to the pressure at point
(3), thus we can “jump across” and then move up to level z2. These steps can be
expressed as
∣
∣ ∣
∣
pA + γ1 zA − z1 − γ2 z1 − z2 = p2 ≈ patm
(2.23)
Another physical reason that we can “jump across” at point (2) is that a continuous length of the same fluid connects these two equal elevations. The hydrostatic
relation (2.14) requires this equality as a form of Pascal’s law:
Any two points at the same elevation in a continuous mass of the same static
fluid will be at the same pressure.
Note that we could not “jump across” from point (1) to a point at the same
elevation in the right-hand tube since these would not be the same elevations
within the same continuous mass of fluid. This idea of jumping across to equal
pressures facilitates multiple-fluid problems. It will be inaccurate, however, if
there are bubbles in the fluid.
Open, pa
zA, pA
Fig. 2.9 Simple open manometer
for measuring pA relative to atmospheric pressure.
A
ρ1
(2)
z1, p1
z 2 , p2 ≈ pa
(1)
Jump across
(3)
p = p1 at z = z1 in fluid 2
ρ2
2.4 Application to Manometry 73
EXAMPLE 2.3
The classic use of a manometer is when two U-tube legs are of equal length, as in
Fig. E2.3, and the measurement involves a pressure difference across two horizontal
points. The typical application is to measure pressure change across a flow device, as
shown. Derive a formula for the pressure difference pa – pb in terms of the system
parameters in Fig. E2.3.
Flow device
(a)
(b)
L
ρ1
h
ρ2
E2.3
Solution
Using Eq. (2.14), start at (a), evaluate pressure changes around the U-tube, and end up
at (b):
pa + ρ1gL + ρ1gh − ρ2gh − ρ1gL = pb
or
pa − pb = (ρ2 − ρ1 )gh
Ans.
The measurement only includes h, the manometer reading. Terms involving L drop out.
Note the appearance of the difference in densities between manometer fluid and working fluid. It is a common student error to fail to subtract out the working fluid density
ρ1—a serious error if both fluids are liquids and less disastrous numerically if fluid 1
is a gas. Academically, of course, such an error is always considered serious by fluid
mechanics instructors.
Although Example 2.3, because of its popularity in engineering experiments,
is sometimes considered to be the “manometer formula,” it is best not to memorize it but rather to adapt Eq. (2.14) to each new multiple-fluid hydrostatics
problem. For example, Fig. 2.10 illustrates a multiple-fluid manometer problem
for finding the difference in pressure between two chambers A and B. By repeating the calculations for manometer fluids with different specific weights, we start
at A, goes down to 1, jumps across, goes up to 2, jumps across, goes down to 3,
jumps across, and finally goes up to B. Thus, in equation form,
pA + γ1(zA − z1) − γ2(z2 − z1) + γ3(z2 − z3) − γ4(zB − z3) = pB
Or the pressure difference between chambers A and B is
pA − pB = −γ1(zA − z1) + γ2(z2 − z1) − γ3(z2 − z3) + γ4(zB − z3)
(2.24)
74
Chapter 2 Pressure Distribution in a Fluid
ρ3
Jump across
z 2, p2
ρ1
zA, pA A
B
Jump across
z1, p1
Fig. 2.10 A complicated multiplefluid manometer to relate pA to
pB. This system is not especially
practical but makes a good
homework or examination problem.
z 2, p2
z1, p1
z 3, p3
Jump across
ρ2
zB, pB
z 3, p3
ρ4
EXAMPLE 2.4
Pressure gage B is to measure the pressure at point A in a water flow. If the pressure at B
is 87 kPa, estimate the pressure at A in kPa. Assume all fluids are at 20˚C. See Fig. E2.4.
SAE 30 oil
Gage B
6 cm
Mercury
A
5 cm
Water
flow
11 cm
4 cm
E2.4
Solution
∙
∙
∙
∙
System sketch: The system is shown in Fig. E2.4.
Assumptions: Hydrostatic fluids, no mixing, vertical “up” in Fig. E2.4.
Approach: Sequential use of Eq. (2.14) to go from A to B.
Property values: From Table 2.1 or Table A.3:
γwater = 9790 N/m3;
γmercury = 133,100 N/m3;
γoil = 8720 N/m3
∙ Solution steps: Proceed from A to B, “down” then “up,” jumping across at the left
mercury meniscus:
∣ ∣
pA + ρw Δz
3
w
∣ ∣ ∣ ∣
− γm Δzm − γo
Δz
o
= pB
3
or pA + (9790 N/m ) (0.05 m) − (133,100 N/m ) (0.07 m) − (8720 N/m3 ) (0.06 m) = 87,000
or pA + 490 − 9317 − 523 = 87,000 Solve for pA = 96,350 N/m2 ≈ 96.4 kPa
2
Ans.
∙ Comments: Note that we abbreviated the units N/m to pascals, or Pa. The intermediate five-figure result, pA = 96,350 Pa, is unrealistic, since the data are known to
only about three significant figures.
2.5 Hydrostatic Forces on Plane Surfaces 75
In making these manometer calculations, we have neglected the capillary height
changes due to surface tension, which were discussed in Example 1.8. These
effects cancel if there is a fluid interface, or meniscus, between similar fluids on
both sides of the U-tube. Otherwise, as in the right-hand U-tube of Fig. 2.10, a
capillary correction can be made or the effect can be made negligible by using
large-bore (≥1 cm) tubes.
2.5 Hydrostatic Forces on Plane Surfaces
The design of containment structures requires computation of the hydrostatic
forces on various solid surfaces adjacent to the fluid. These forces relate to the
weight of fluid bearing on the surface. For example, a container with a flat,
horizontal bottom of area Ab and water depth H will experience a downward bottom force Fb = γHAb. If the surface is not horizontal, additional computations are
needed to find the horizontal components of the hydrostatic force.
If we neglect density changes in the fluid, Eq. (2.14) applies and the pressure on any submerged surface varies linearly with depth. For a plane surface,
the linear stress distribution is exactly analogous to combined bending and
compression of a beam in strength-of-materials theory. The hydrostatic problem thus reduces to simple formulas involving the centroid and moments of
inertia of the plate cross-sectional area.
Figure 2.11 shows a plane panel of arbitrary shape completely submerged in
a liquid. The panel plane makes an arbitrary angle θ with the horizontal free
surface, so that the depth varies over the panel surface. If h is the depth to any
element area dA of the plate, from Eq. (2.14) the pressure there is p = pa + γh.
Free surface
p = pa
θ
h (x, y)
Resultant
force:
F = pCG A
hCG
ξ=
y
Side view
CG
Fig. 2.11 Hydrostatic force and
center of pressure on an arbitrary
plane surface of area A inclined at
an angle θ below the free surface.
x
dA = dx dy
CP
Plan view of arbitrary plane surface
h
sin θ
76
Chapter 2 Pressure Distribution in a Fluid
To derive formulas involving the plate shape, establish an xy coordinate system
in the plane of the plate with the origin at its centroid CG shown in Fig. 2.11,
plus a dummy coordinate ξ down from the surface in the plane of the plate. Then
the total hydrostatic force on one side of the plate is given by
∫
∫
∫
F = p dA = ( pa + γh) dA = pa A + γ h dA
(2.25)
The remaining integral is evaluated by noticing from Fig. 2.11 that h = ξ sin
θ and, by definition, the centroidal slant distance from the surface to the plate is
ξCG =
∫
1
ξ dA
A
Therefore, since θ is constant along the plate, Eq. (2.25) becomes
∫
F = pa A + γ sin θ ξ dA = pa A + γ sin θ ξCG A
Finally, unravel this by noticing that ξCG sin θ = hCG, the depth straight down
from the surface to the plate centroid. Thus
F = paA + γhCG A = ( pa + γhCG )A = pCG A (2.26)
The force on one side of any plane submerged surface in a uniform fluid equals
the pressure at the plate centroid times the plate area, independent of the shape
of the plate or the angle θ at which it is slanted.
Equation (2.26) can be visualized physically in Fig. 2.12 as the resultant of a
linear stress distribution over the plate area. This simulates combined compression
p = pa
Free surface
θ
Line of action
Resultant
force:
F = pCG A
Fig. 2.12 The hydrostatic pressure
force on a plane surface is equal,
regardless of its shape, to the
resultant linear pressure distribution
on that surface F = pCGA. Its line of
action passes through the center of
pressure (CP) where yCP is located
below the centroid, xCP is located
either to the left or right of the
centroid.
hCG
ξCG
yCP
Side view
y
CG
xCP
x
CP
Plan view of arbitrary plane surface
2.5 Hydrostatic Forces on Plane Surfaces 77
and bending of a beam of the same cross section. It follows that the “bending”
portion of the stress causes no force if its “neutral axis” passes through the plate
centroid of area. Thus the remaining “compression” part must equal the centroid
stress times the plate area. This is the result of Eq. (2.26).
However, to balance the bending-moment portion of the stress, the resultant
force F acts not through the centroid but below it toward the high-pressure side.
Its line of action passes through the center of pressure CP of the plate, as
sketched in Fig. 2.12. To find the coordinates (xCP, yCP), we sum moments of
the elemental force p dA about the centroid and equate to the moment of the
resultant F. To compute yCP, we equate
∫
∫
∫
FyCP = yp dA = y( pa + γξ sin θ) dA = γ sin θ yξ dA
The term ∫ pay dA vanishes by definition of centroidal axes. Introducing ξ = ξCG
– y, we obtain
FyCP = γ sin θ (ξCG y dA − y2 dA) = −γ sin θ Ixx
∫
∫
where again ∫ y dA = 0 and Ixx is the area moment of inertia of the plate area
about its centroidal x axis, computed in the plane of the plate. Substituting for F
gives the result
yCP = −γ sin θ
Ixx
pCG A
(2.27)
The negative sign in Eq. (2.27) shows that yCP is below the centroid at a deeper
level and, unlike F, depends on angle θ. If we move the plate deeper, yCP
approaches the centroid because every term in Eq. (2.27) remains constant except
pCG, which increases.
The determination of xCP is exactly similar:
∫
∫
FxCP = xp dA = x[pa + γ(ξCG − y) sin θ] dA
∫
= −γ sin θ xy dA = −γ sin θ Ixy
where Ixy is the product of inertia of the plate, again computed in the plane of
the plate. Substituting for F gives
xCP = −γ sin θ
Ixy
pCGA
(2.28)
For positive Ixy, xCP is negative because the dominant pressure force acts in the
third, or lower left, quadrant of the panel. If Ixy = 0, usually implying symmetry,
xCP = 0 and the center of pressure lies directly below the centroid on the y axis.
78
Chapter 2 Pressure Distribution in a Fluid
Gage Pressure Formulas
In most cases the ambient pressure pa is neglected because it acts on both sides
of the plate; for example, the other side of the plate is inside a ship or on the dry
side of a gate or dam. In this case pCG = γhCG, and the center of pressure becomes
independent of specific weight:
F = γhCGA
yCP = −
Ixx sin θ
hCG A
xCP = −
Ixy sin θ
hCG A
(2.29)
Figure 2.13 gives the area and moments of inertia of several common cross sections for use with these formulas. Note that θ is the angle between the plate and
the horizon.
L
2
y
A = bL
x
Ixx =
L
2
b
2
(a)
A = π R2
y
bL3
12
x
R
Ix y = 0
b
2
Ixx =
R
πR 4
4
Ixy = 0
(b)
s
A = bL
2
2L
3
y
Ixx =
x
L
3
Ix y =
b
2
b
2
A=
Ixx = 0.10976R 4
bL3
36
Ix y = 0
y
2s)L 2
b(b –
72
x
R
(c)
πR2
4
Ixx = 0.05488R 4
y
Ix y = 0.01647R 4
x
x
Fig. 2.13 Centroidal moments of
inertia for various cross sections:
(a) rectangle, (b) circle, (c) triangle,
(d) semicircle, (e) quarter circle, and
(f ) semi-parabola.
R
4R
3π
(e)
a
3a
5
3b
8
( f)
2ab
3
8ba 3
Ixx =
175
3
Ix y = 19b a
480
A=
b
A=
4R
3π
4R
3π
R
(d)
y
πR2
2
2.5 Hydrostatic Forces on Plane Surfaces 79
EXAMPLE 2.5
The gate in Fig. E2.5a is 5 ft wide, is hinged at point B, and rests against a smooth
wall at point A. Compute (a) the force on the gate due to seawater pressure, (b) the
horizontal force P exerted by the wall at point A, and (c) the reactions at the hinge B.
Wall
pa
Seawater:
64 lbf/ft 3
15 ft
A
pa
Gate
6 ft
B
E2.5a
θ
8 ft
Hinge
Solution
Part (a)
By geometry the gate is 10 ft long from A to B, and its centroid is halfway between, or at elevation 3 ft above point B. The depth hCG is thus 15 − 3 = 12 ft. The gate
area is 5(10) = 50 ft2. Neglect pa as acting on both sides of the gate. From Eq. (2.26) the
hydrostatic force on the gate is
F = pCG A = γhCG A = (64 lbf/ft3 ) (12 ft) (50 ft2 ) = 38,400 lbf
Ans. (a)
Part (b)
First we must find the center of pressure of F. A free-body diagram of the gate is
shown in Fig. E2.5b. The gate is a rectangle, hence
Ixy = 0 and Ixx =
bL3 (5 ft) (10 ft) 3
=
= 417 ft4
12
12
The distance l from the CG to the CP is given by Eqs. (2.29) since pa is neglected.
l = −yCP = +
(417 ft4 ) ( 106 )
Ixx sin θ
=
= 0.417 ft
hCG A
(12 ft) (50 ft2 )
A
F
5 ft
l
Bx
E2.5b
B
Bz
θ
CP
CG
L = 10 ft
P
80
Chapter 2 Pressure Distribution in a Fluid
The distance from point B to force F is thus 10 − l − 5 = 4.583 ft. Summing the
moments counterclockwise about B gives
PL sin θ − F(5 − l ) = P(6 ft) − (38,400 lbf ) (4.583 ft) = 0
or
P = 29,300 lbf
Ans. (b)
Part (c)
With F and P known, the reactions Bx and Bz are found by summing forces on the gate:
g Fx = 0 = Bx + F sin θ − P = Bx + 38,400 lbf (0.6) − 29,300 lbf
or
Bx = 6300 lbf
g Fz = 0 = Bz − F cos θ = Bz − 38,400 lbf (0.8)
or
Bz = 30,700 lbf
Ans. (c)
This example should have reviewed your knowledge of statics.
The solution of Example 2.5 was achieved with the moment of inertia
formulas, Eqs. (2.29). They simplify the calculations, but one loses a physical
feeling for the forces. Let us repeat Parts (a) and (b) of Example 2.5 using a
more visual approach.
EXAMPLE 2.6
Repeat Example 2.5 to sketch the pressure distribution on plate AB, and break this
distribution into rectangular and triangular parts to solve for (a) the force on the plate
and (b) the center of pressure.
Solution
Part (a)
Point A is 9 ft deep, hence pA = γhA = (64 lbf/ft3)(9 ft) = 576 lbf/ft2. Similarly, Point
B is 15 ft deep, hence pB = γhB = (64 lbf/ft3)(15 ft) = 960 lbf/ft2. This defines the
linear pressure distribution in Fig. E2.6. The rectangle is 576 by 10 ft by 5 ft into the
paper. The triangle is (960 − 576) = 384 lbf/ft2 × 10 ft by 5 ft. The centroid of the
rectangle is 5 ft down the plate from A. The centroid of the triangle is 6.67 ft down
from A. The total force is the rectangle force plus the triangle force:
lbf
384 lbf
F = (576 2 ) (10 ft) (5 ft) + (
(10 ft) (5 ft)
2 ft2 )
ft
= 28,800 lbf + 9600 lbf = 38,400 lbf
576 lbf/ft2
F
960
t
5f
lbf/ft2
l
5-
E2.6
B
l
A
6 ft
8 ft
Ans. (a)
2.5 Hydrostatic Forces on Plane Surfaces 81
Part (b)
The moments of these forces about point A are
ΣMA = (28,800 lbf ) (5 ft) + (9600 lbf ) (6.67 ft) = 144,000 + 64,000 = 208,000 ft · lbf
Then
5 ft + l =
MA 208,000 ft · lbf
=
= 5.417 ft hence l = 0.417 ft
F
38,400 lbf
Ans. (b)
Comment: We obtain the same force and center of pressure as in Example 2.5 but
with more understanding. However, this approach is awkward and laborious if the plate
is not a rectangle. It would be difficult to solve Example 2.7 with the pressure distribution alone because the plate is a triangle. Thus moments of inertia can be a useful
simplification.
EXAMPLE 2.7
A tank of oil has a right-triangular panel near the bottom, as in Fig. E2.7. Omitting
pa, find the (a) hydrostatic force and (b) CP on the panel.
pa
Oil: ρ = 800 kg/m 3
5m
30°
11 m
4m
6m
pa
CG
CP
4m
8m
4m
E2.7
2m
Solution
Part (a)
The triangle has properties given in Fig. 2.13c. The centroid is one-third up (4 m) and
­one-third over (2 m) from the lower left corner, as shown. The area is
1
2 (6
m) (12 m) = 36 m2
The moments of inertia are
and
Ixx =
bL3 (6 m) (12 m) 3
=
= 288 m4
36
36
Ixy =
b(b − 2s)L2 (6 m) [6 m − 2(6 m) ] (12 m) 2
=
= −72 m4
72
72
82
Chapter 2 Pressure Distribution in a Fluid
The depth to the centroid is hCG = 5 + 4 = 9 m; thus the hydrostatic force from
Eq. (2.26) is
F = ρghCG A = (800 kg /m3 ) (9.807 m /s2 ) (9 m) (36 m2 )
= 2.54 × 106 (kg · m)/s2 = 2.54 × 106 N = 2.54 MN
Ans. (a)
Part (b)
The CP position is given by Eqs. (2.29):
yCP = −
xCP = −
Ixx sin θ
(288 m4 ) (sin 30°)
=−
= −0.444 m
hCG A
(9 m) (36 m2 )
Ixy sin θ
hCG A
=−
(−72 m4 ) (sin 30°)
(9 m) (36 m2 )
= +0.111 m
Ans. (b)
The resultant force F = 2.54 MN acts through this point, which is down and to the
right of the centroid, as shown in Fig. E2.7.
2.6 Hydrostatic Forces on Curved Surfaces
The resultant pressure force on a curved surface is most easily computed by
separating it into horizontal and vertical components. Consider the arbitrary
curved surface sketched in Fig. 2.14a. The incremental pressure forces, being
normal to the local area element, vary in direction along the surface and thus
cannot be added numerically. We could sum the separate three components of
these elemental pressure forces, but it turns out that we need not perform a laborious three-way integration.
Figure 2.14b shows a free-body diagram of the column of fluid contained in
the vertical projection above the curved surface. The desired forces FH and FV
are exerted by the surface on the fluid column. Other forces are shown due to
fluid weight and horizontal pressure on the vertical sides of this column. The
column of fluid must be in static equilibrium. On the upper part of the column
bcde, the horizontal components F1 exactly balance and are not relevant to the
Wair
d
Curved surface
projection onto
vertical plane
FV
Fig. 2.14 Computation of
hydrostatic force on a curved
surface: (a) submerged curved
surface; (b) free-body diagram of
fluid above the curved surface.
FH
FH
F1
F1
W1
c
b
W2
FH
a
FV
(a)
e
(b)
FH
2.6 Hydrostatic Forces on Curved Surfaces 83
discussion. On the lower, irregular portion of fluid abc adjoining the surface,
summation of horizontal forces shows that the desired force FH due to the curved
surface is exactly equal to the force FH on the vertical left side of the fluid column. This left-side force can be computed by the plane surface formula, Eq.
(2.26), based on a vertical projection of the area of the curved surface. This is a
general rule and simplifies the analysis:
The horizontal component of force on a curved surface equals the force on the
plane area formed by the projection of the curved surface onto a vertical plane
normal to the component.
If there are two horizontal components, both can be computed by this scheme.
Summation of vertical forces on the fluid free body then shows that
FV = W1 + W2 + Wair
(2.30)
We can state this in words as our second general rule:
The vertical component of pressure force on a curved surface equals in magnitude and direction the weight of the entire column of fluid, both liquid and
atmosphere, above the curved surface.
Thus the calculation of FV involves little more than finding centers of mass of a
column of fluid—perhaps a little integration if the lower portion abc in Fig. 2.14b
has a particularly vexing shape.
EXAMPLE 2.8
A dam has a parabolic shape z/z0 = (x/x0)2 as shown in Fig. E2.8a, with x0 = 10 ft and
z0 = 24 ft. The fluid is water, γ = 62.4 lbf/ft3, and atmospheric pressure may be omitted. Compute the forces FH and FV on the dam and their line of action. The width of
the dam is 50 ft.
pa = 0 lbf/ft2 gage
FV
z
z0
FH
x
x0
E2.8a
( (
x
z = z0 x
0
2
Solution
∙ System sketch: Figure E2.8b shows the various dimensions. The dam width is b = 50 ft.
∙ Approach: Calculate FH and its line of action from Eqs. (2.26) and (2.29). Calculate
FV and its line of action by finding the weight of fluid above the parabola and the
centroid of this weight.
84
Chapter 2 Pressure Distribution in a Fluid
∙ Solution steps for the horizontal component: The vertical projection of the parabola
lies along the z axis in Fig. E2.8b and is a rectangle 24 ft high and 50 ft wide. Its
centroid is halfway down, or hCG = 24/2 = 12 ft. Its area is Aproj = (24 ft)(50 ft) =
1200 ft2. Then, from Eq. (2.26),
lbf
FH = γ hCG Aproj = (62.4 3 ) (12 ft) (1200 ft2 ) = 898,560 lbf ≈ 899 × 103 lbf
ft
The line of action of FH is below the centroid of Aproj, as given by Eq. (2.29):
yCP, proj = −
Ixx sin θ
(1/12) (50 ft) (24 ft) 3 sin 90°
=−
= −4 ft
hCG Aproj
(12 ft) (1200 ft2 )
Thus FH is 12 + 4 = 16 ft, or two-thirds of the way down from the surface (8 ft up
from the bottom).
∙ Comments: Note that you calculate FH and its line of action from the vertical projection of the parabola, not from the parabola itself. Since this projection is vertical,
its angle θ = 90˚.
∙ Solution steps for the vertical component: The vertical force FV equals the weight
of water above the parabola. Alas, a parabolic section is not in Fig. 2.13, so we had
to look it up in another book. The area and centroid are shown in Fig. E2.8b. The
weight of this parabolic amount of water is
FV = γ Asectionb = (62.4
lbf 2
(24 ft) (10 ft) ] (50 ft) = 499,200 lbf ≈ 499 × 103 lbf
ft3 )[ 3
z0 = 24 ft
3z0
5
Area =
2 x0z 0
3
FV
Parabola
0
E2.8b
3x 0
8
x0 = 10 ft
This force acts downward, through the centroid of the parabolic section, or at a
distance 3x0/8 = 3.75 ft over from the origin, as shown in Figs. E2.8b,c. The resultant hydrostatic force on the dam is
F = (FH2 + FV2 ) 1/2 = [ (899E3 lbf) 2 + (499E3 lbf) 2 ] 1/2 = 1028 × 103 lbf at
29° Ans.
This resultant is shown in Fig. E2.8c and passes through a point 8 ft up and 3.75 ft
over from the origin. It strikes the dam at a point 5.43 ft over and 7.07 ft up, as shown.
∙ Comments: Note that entirely different formulas are used to calculate FH and FV.
The concept of center of pressure CP is, in the writer’s opinion, stretched too far
when applied to curved surfaces.
2.7 Hydrostatic Forces in Layered Fluids 85
z
Resultant = 1028 × 103 lbf acts along z = 10.083 – 0.5555x
3.75 ft
FV = 499 × 103 lbf
FH = 899 × 103 lbf
29°
Parabola z = 0.24x2
7.07 ft
8 ft
E2.8c
B
A
5.43 ft
x
EXAMPLE 2.9
Find an algebraic formula for the net vertical force F on the submerged semicircular
projecting structure CDE in Fig. E2.9. The structure has uniform width b into the paper.
The liquid has specific weight γ.
FU
C
Solution
R
E
FL
E2.9
0
D
The net force is the difference between the upward force FL on the lower surface DE
and the downward force FU on the upper surface CD, as shown in Fig. E2.9. The force
FU equals γ times the volume ABDC above surface CD. The force FL equals γ times the
volume ABDEC above surface DE. The latter is clearly larger. The difference is γ times
the volume of the structure itself. Thus the net upward fluid force on the semicylinder is
F = γfluid (volume CDE ) = γfluid
π 2
R b
2
Ans.
This is the principle upon which the laws of buoyancy, Sec. 2.8, are founded. Note that
the result is independent of the depth of the structure and depends upon the specific
weight of the fluid, not the material within the structure.
2.7 Hydrostatic Forces in Layered Fluids
The formulas for plane and curved surfaces in Secs. 2.5 and 2.6 are valid only
for a fluid of uniform density. If the fluid is layered with different densities, as
in Fig. 2.15, a single formula cannot solve the problem because the slope of the
linear pressure distribution changes between layers. However, the formulas apply
separately to each layer, and thus the appropriate remedy is to compute and sum
the separate layer forces and moments.
Consider the slanted plane surface immersed in a two-layer fluid in Fig. 2.15.
The slope of the pressure distribution becomes steeper as we move down into the
denser second layer. The total force on the plate does not equal the pressure at
86
Chapter 2 Pressure Distribution in a Fluid
z
F1 = p
A1
CG1
Plane
surface
z=0
pa
ρ1 < ρ2
Fluid 1
p = pa – ρ1gz
z 1, p1
p1 = pa – ρ1gz1
F2 = pCG A 2
2
ρ
2
Fluid 2
z 2 , p2
p = p1 – ρ2 g(z – z 1)
Fig. 2.15 Hydrostatic forces on a
surface immersed in a layered fluid
must be summed in separate pieces.
p2 = p1 – ρ 2 g(z 2 – z 1)
the centroid times the plate area, but the plate portion in each layer does satisfy
the formula, so that we can sum forces to find the total:
F = g Fi = g pCGiAi
(2.31)
Similarly, the centroid of the plate portion in each layer can be used to locate the
center of pressure on that portion:
yCPi = −
ρig sin θi Ixxi
pCGi Ai
xCPi = −
ρig sin θi Ixyi
pCGi Ai
(2.32)
These formulas locate the center of pressure of that particular Fi with respect to
the centroid of that particular portion of plate in the layer, not with respect to the
centroid of the entire plate. The center of pressure of the total force F = Σ Fi can
then be found by summing moments about some convenient point such as the
surface. The following example will illustrate this.
EXAMPLE 2.10
A tank 20 ft deep and 7 ft wide is layered with 8 ft of oil, 6 ft of water, and 4 ft of
mercury. Compute (a) the total hydrostatic force and (b) the resultant center of pressure
of the fluid on the right-hand side of the tank.
Solution
Part (a)
Divide the end panel into three parts as sketched in Fig. E2.10, and find the hydrostatic
­pressure at the centroid of each part, using the relation (2.26) in steps as in Fig. E2.10:
2.7 Hydrostatic Forces in Layered Fluids 87
pCG1 = (55.0 lbf/ft3 ) (4 ft) = 220 lbf/ft2
pCG2 = (55.0) (8) + 62.4(3) = 627 lbf/ft2
pCG3 = (55.0) (8) + 62.4(6) + 846(2) = 2506 lbf/ft2
These pressures are then multiplied by the respective panel areas to find the force on
each portion:
F1 = pCG1A1 = (220 lbf/ft2 ) (8 ft) (7 ft) = 12,300 lbf
F2 = pCG2A2 = 627(6) (7) = 26,300 lbf
F3 = pCG3A3 = 2506(4) (7) = 70,200 lbf
F = g Fi = 108,800 lbf
pa = 0
Oi
l: 5
5.0
Wa
ter
Me
rcu
z=0
7 ft
4 ft
(1)
11 ft
lbf
/ft 3
8 ft
(62
.4)
ry
Ans. (a)
6 ft
(84
6)
16 ft
(2)
4 ft (3)
E2.10
Part (b)
Equations (2.32) can be used to locate the CP of each force Fi, noting that θ = 90˚ and sin θ = 1
for all parts. The moments of inertia are Ixx1 = (7 ft)(8 ft)3/12 = 298.7 ft4, Ixx2 = 7(6)3/12
= 126.0 ft4, and Ixx3 = 7(4)3/12 = 37.3 ft4. The centers of pressure are thus at
yCP1 = −
yCP2
ρ1gIxx1
F1
=−
(55.0 lbf/ft3 ) (298.7 ft4 )
= −1.33 ft
12,300 lbf
62.4(126.0)
=−
= −0.30 ft
26,300
yCP3 = −
846(37.3)
= −0.45 ft
70,200
This locates zCP1 = −4 − 1.33 = −5.33 ft, zCP2 = −11 − 0.30 = −11.30 ft, and zCP3 = −16 −
0.45 = −16.45 ft. Summing moments about the surface then gives
gFizCPi = FzCP
or
or
12,300(−5.33) + 26,300(−11.30) + 70,200(−16.45) = 108,800zCP
zCP = −
1,518,000
= −13.95 ft
108,800
Ans. (b)
The center of pressure of the total resultant force on the right side of the tank lies 13.95
ft below the surface.
88
Chapter 2 Pressure Distribution in a Fluid
2.8 Buoyancy and Stability
The same principles used to compute hydrostatic forces on surfaces can be applied
to the net pressure force on a completely submerged or floating body. The results
are the two laws of buoyancy discovered by Archimedes in the third century b.c.:
1. A body immersed in a fluid experiences a vertical buoyant force equal to
the weight of the fluid it displaces.
2. A floating body displaces its own weight in the fluid in which it floats.
Archimedes (287–212 b.c.) was born and lived in the Greek city-state of Syracuse, on what is now the island of Sicily. He was a brilliant mathematician and
engineer, two millennia ahead of his time. He calculated an accurate value for pi
and approximated areas and volumes of various bodies by summing elemental
shapes. In other words, he invented the integral calculus. He developed levers,
pulleys, catapults, and a screw pump. Archimedes was the first to write large
numbers as powers of 10, avoiding Roman numerals. And he deduced the principles of buoyancy, which we study here, when he realized how light he was
when sitting in a bathtub.
Archimedes’ two laws are easily derived by referring to Fig. 2.16. In Fig. 2.16a,
the body lies between an upper curved surface 1 and a lower curved surface 2.
From Eq. (2.30) for vertical force, the body experiences a net upward force
FB = FV (2) − FV (1)
= (fluid weight above 2) − (fluid weight above 1)
= weight of fluid equivalent to body volume
(2.33)
Alternatively, from Fig. 2.16b, we can sum the vertical forces on elemental vertical slices through the immersed body:
FB =
∫
body
FV (1)
∫
(p2 − p1 ) dAH = −γ (z2 − z1 ) dAH = (γ)(body volume)
Surface
1
p1
Horizontal
elemental
area d AH
z1 – z 2
Fig. 2.16 Two different approaches
to the buoyant force on an arbitrary
immersed body: (a) forces on upper
and lower curved surfaces;
(b) summation of elemental
vertical-pressure forces.
Surface
2
FV (2)
(a)
p2
(b)
(2.34)
2.8 Buoyancy and Stability 89
Neglect the displaced air up here.
CG
W
B
Fig. 2.17 Static equilibrium of a
floating body.
FB
(Displaced volume) × (γ of fluid) = body weight
These are identical results and equivalent to Archimedes’ law 1.
Equation (2.34) assumes that the fluid has uniform specific weight. The line of
action of the buoyant force passes through the center of volume of the displaced
body; that is, its center of mass computed as if it had uniform density. This point
through which FB acts is called the center of buoyancy, commonly labeled B or
CB on a drawing. Of course, the point B may or may not correspond to the actual
center of mass of the body’s own material, which may have variable density.
Equation (2.34) can be generalized to a layered fluid (LF) by summing the
weights of each layer of density ρi displaced by the immersed body:
(FB ) LF = g ρig(displaced volume) i (2.35)
Each displaced layer would have its own center of volume, and one would have
to sum moments of the incremental buoyant forces to find the center of buoyancy
of the immersed body.
Since liquids are relatively heavy, we are conscious of their buoyant forces, but
gases also exert buoyancy on any body immersed in them. For example, human
beings have an average specific weight of about 60 lbf/ft3. We may record the
weight of a person as 180 lbf and thus estimate the person’s total volume as 3.0
ft3. However, in so doing we are neglecting the buoyant force of the air surrounding the person. At standard conditions, the specific weight of air is 0.0763 lbf/ft3;
hence the buoyant force is approximately 0.23 lbf. If measured in a vacuum, the
person would weigh about 0.23 lbf more. For balloons and blimps the buoyant
force of air, instead of being negligible, is the controlling factor in the design.
Also, many flow phenomena, such as natural convection of heat and vertical mixing in the ocean, are strongly dependent on seemingly small buoyant forces.
Floating bodies are a special case; only a portion of the body is submerged,
with the remainder poking up out of the free surface. This is illustrated in Fig. 2.17,
where the shaded portion is the displaced volume. Equation (2.34) is modified to
apply to this smaller volume:
FB = (γ)(displaced volume) = floating-body weight
(2.36)
Not only does the buoyant force equal the body weight, but also they are collinear since there can be no net moments for static equilibrium. Equation (2.36)
is the mathematical equivalent of Archimedes’ law 2 stated at the beginning of
this section.
90
Chapter 2 Pressure Distribution in a Fluid
EXAMPLE 2.11
A block of concrete weighs 100 lbf in air and “weighs” only 60 lbf when immersed in
fresh water (62.4 lbf/ft3). What is the average specific weight of the block?
Solution
60 lbf
A free-body diagram of the submerged block (see Fig. E2.11) shows a balance between
the apparent weight, the buoyant force, and the actual weight:
FB
or
W = 100 lbf
g Fz = 0 = 60 + FB − 100
FB = 40 lbf = (62.4 lbf/ft3 ) (block volume, ft3 )
Solving gives the volume of the block as 40/62.4 = 0.641 ft3. Therefore, the specific
weight of the block is
γblock =
E2.11
100 lbf
=156 lbf/ft3 0.641 ft3
Ans.
Occasionally, a body will have exactly the right weight and volume for its ratio
to equal the specific weight of the fluid. If so, the body will be neutrally buoyant
and will remain at rest at any point where it is immersed in the fluid. Small,
neutrally buoyant particles are sometimes used in flow visualization, and a neutrally buoyant body called a Swallow float [2] is used to track oceanographic
currents. A submarine can achieve positive, neutral, or negative buoyancy by
pumping water into or out of its ballast tanks.
Stability
A floating body as in Fig. 2.17 may not approve of the position in which it is
floating. If so, it will overturn at the first opportunity and is said to be statically
unstable, like a pencil balanced on its point. The least disturbance will cause it
to seek another equilibrium position that is stable. Engineers must design to avoid
floating instability. The only way to tell for sure whether a floating position is
stable is to “disturb” the body a slight amount mathematically and see whether
it develops a restoring moment that will return it to its original position. If so, it
is stable; if not, unstable. Such calculations for arbitrary floating bodies have been
honed to a fine art by naval architects [3], but we can at least outline the basic
principle of the static stability calculation. Figure 2.18 illustrates the computation
for the usual case of a symmetric floating body. The steps are as follows:
1. The basic floating position is calculated from Eq. (2.36). The body’s center
of mass G and center of buoyancy B are computed.
2. The body is tilted a small angle Δθ, and a new waterline is established
for the body to float at this angle. The new position B′ of the center of
buoyancy is calculated. A vertical line drawn upward from B′ intersects
the line of symmetry at a point M, called the metacenter, which is independent of Δθ for small angles.
2.8 Buoyancy and Stability 91
Small
∆θ disturbance
angle
Line of
symmetry
Small
disturbance
angle
∆θ
M
Fig. 2.18 Calculation of the
metacenter M of the floating body
shown in (a). Tilt the body a small
angle Δθ. Either (b) B′ moves far
out (point M above G denotes
stability); or (c) B′ moves slightly
(point M below G denotes
instability).
FB
W
FB
W
W
Either
M
FB
B'
B
(a)
G
G
G
B'
Restoring moment
(b)
or
Overturning moment
(c)
3. If point M is above G (that is, if the metacentric height MG is positive), a
restoring moment is present and the original position is stable. If M is
below G (negative MG), the body is unstable and will overturn if disturbed. Stability increases with increasing MG.
Thus the metacentric height is a property of the cross section for the given weight,
and its value gives an indication of the stability of the body. For a body of varying cross section and draft, such as a ship, the computation of the metacenter can
be very involved.
Stability Related to Waterline Area1
Naval architects [3] have developed the general stability concepts from Fig. 2.18
into a simple computation involving the area moment of inertia of the waterline area (as seen from above) about the axis of tilt. The derivation—see [3] for
details—assumes that the body has a smooth shape variation (no discontinuities)
near the waterline. Recall that M is the metacenter, B is the center of buoyancy,
and G is the center of gravity. The final elegant formula relates the distances
between these points:
MG =
IO
− GB
vsub
(2.37)
Where IO is the area moment of inertia of the waterline area about the tilt axis
O and vsub is the volume of the submerged portion of the floating body. It is
desirable, of course, that MG be positive for the body to be stable.
The engineer locates G and B from the basic shape and design of the floating
body and then calculates IO and vsub to determine if MG is positive.
Engineering design counts upon effective operation of the results. A stability
analysis is useless if the floating body runs aground on rocks, as in Fig. 2.19.
1
This section may be omitted without loss of continuity.
92
Chapter 2 Pressure Distribution in a Fluid
Fig. 2.19 The Italian liner Costa
Concordia aground on January 14,
2012. Stability analysis may fail
when operator mistakes occur
(Gregorio Borgia/AP Images).
EXAMPLE 2.12
A barge has a uniform rectangular cross section of width 2L and vertical draft of height
H, as in Fig. E2.12. Determine (a) the metacentric height for a small tilt angle and (b)
the range of ratio L/H for which the barge is statically stable if G is exactly at the
waterline as shown.
G
O
L
E2.12
H
●B
L
Solution
If the barge has length b into the paper, the waterline area, relative to tilt axis O, has a
base b and a height 2L; therefore, IO = b(2L)3/12. Meanwhile, υsub = 2LbH. Equation
(2.37) predicts
MG =
Io
8bL3/12 H
L2
H
− GB =
− =
− υsub
2LbH
2 3H 2
Ans. (a)
The barge can thus be stable only if
L2 > 3H2/2 or 2L > 2.45H Ans. (b)
The wider the barge relative to its draft, the more stable it is. Lowering G would help also.
2.9 Pressure Distribution in Rigid-Body Motion 93
Fig. 2.20 A North Atlantic iceberg
formed by calving from a Greenland
glacier. These, and their even larger
Antarctic sisters, are the largest
floating bodies in the world.
(Dmitry Reznichenko/Shutterstock)
Even an expert will have difficulty determining the floating stability of a buoyant body of irregular shape. Such bodies may have two or more stable positions.
For example, a ship may float the way we like it, so that we can sit on the deck,
or it may float upside down (capsized). An interesting mathematical approach to
floating stability is given in Ref. 4. The author of this reference points out that
even simple shapes, such as a cube of uniform density, may have a great many
stable floating orientations, not necessarily symmetric. Homogeneous circular
cylinders can float with the axis of symmetry tilted from the vertical.
Floating instability occurs in nature. Fish generally swim with their planes of symmetry vertical. After death, this position is unstable and they float with their flat sides
up. Giant icebergs may overturn after becoming unstable when their shapes change
due to underwater melting. Iceberg overturning is a dramatic, rarely seen event.
Figure 2.20 shows a typical North Atlantic iceberg formed by calving from a
Greenland glacier that protruded into the ocean. The exposed surface is rough,
indicating that it has undergone further calving. Icebergs are frozen fresh, bubbly,
glacial water of average density 900 kg/m3. Thus, when an iceberg is floating in
seawater, whose average density is 1025 kg/m3, approximately 900/1025, or
seven-eighths, of its volume lies below the water.
2.9 Pressure Distribution in Rigid-Body Motion
In rigid-body motion, all particles are in combined translation and rotation, and
there is no relative motion between particles. With no relative motion, there are
no strains or strain rates, so that the viscous term in Eq. (2.8) vanishes, leaving
a balance between pressure, gravity, and particle acceleration:
∇p = ρ(g − a)
(2.38)
The pressure gradient acts in the direction g – a, and lines of constant pressure (including the free surface, if any) are perpendicular to this direction. The general case of
combined translation and rotation of a rigid body is discussed in Chap. 3, Fig. 3.11.
94
Chapter 2 Pressure Distribution in a Fluid
z
ax
a
az
x
–a
θ = tan –1
θ
g
Fig. 2.21 Tilting of constant-pressure
surfaces in a tank of liquid in rigidbody acceleration.
∇p ∝g – a
az
ax
S
ax
g + az
Fluid
at rest
p = p1
p2
p3
Fluids can rarely move in rigid-body motion unless restrained by confining
walls for a long time. For example, suppose a tank of water is in a car that starts
a constant acceleration. The water in the tank would begin to slosh about, and that
sloshing would damp out very slowly until finally the particles of water would be
in approximately rigid-body acceleration. This would take so long that the car
would have reached hypersonic speeds. Nevertheless, we can at least discuss the
pressure distribution in a tank of rigidly accelerating water.
Uniform Linear Acceleration
In the case of uniform rigid-body acceleration, Eq. (2.38) applies, a having the
same magnitude and direction for all particles. With reference to Fig. 2.21, the
parallelogram sum of g and –a gives the direction of the pressure gradient or
greatest rate of increase of p. The surfaces of constant pressure must be perpendicular to this and are thus tilted at a downward angle θ such that
ax
θ = tan−1
(2.39)
g + az
One of these tilted lines is the free surface, which is found by the requirement
that the fluid retain its volume unless it spills out. The rate of increase of pressure
in the direction g – a is greater than in ordinary hydrostatics and is given by
dp
= ρG where G = [a2x + (g + az ) 2 ] 1/2
ds
(2.40)
These results are independent of the size or shape of the container as long as the
fluid is continuously connected throughout the container.
EXAMPLE 2.13
A drag racer rests her coffee mug on a horizontal tray while she accelerates at 7 m/s2. The
mug is 10 cm deep and 6 cm in diameter and contains coffee 7 cm deep at rest. (a) Assuming rigid-body acceleration of the coffee, determine whether it will spill out of the mug. (b)
Calculate the gage pressure in the corner at point A if the density of coffee is 1010 kg/m3.
2.9 Pressure Distribution in Rigid-Body Motion 95
Solution
∙ System sketch: Figure E2.13 shows the coffee tilted during the acceleration.
3 cm
∆z
θ
7 cm
ax = 7 m/s2
A
E2.13
3 cm
∙ Assumptions: Rigid-body horizontal acceleration, ax = 7 m/s2. Symmetric coffee
cup.
∙ Property values: Density of coffee given as 1010 kg/m3.
∙ Approach (a): Determine the angle of tilt from the known acceleration, then find
the height rise.
∙ Solution steps: From Eq. (2.39), the angle of tilt is given by
ax
7.0 m/s2
θ = tan−1 g = tan−1
= 35.5°
9.81 m/s2
If the mug is symmetric, the tilted surface will pass through the center point of the rest
position, as shown in Fig. E2.13. Then the rear side of the coffee free surface will rise
an amount Δz given by
Δz = (3 cm) (tan 35.5°) = 2.14 cm < 3 cm therefore no spilling
Ans. (a)
∙ Comment (a): This solution neglects sloshing, which might occur if the start-up is
uneven.
∙ Approach (b): The pressure at A can be computed from Eq. (2.40), using the perpendicular distance Δs from the surface to A. When at rest, pA = ρghrest = (1010 kg/m3)
(9.81 m/s2)(0.07 m) = 694 Pa.
kg
pA = ρG Δs = (1010 3)[√ (9.81) 2 + (7.0) 2] [(0.07 + 0.0214) cos 35.5°] ≈ 906 Pa Ans. (b)
m
∙ Comment (b): The acceleration has increased the pressure at A by 31 percent. Think
about this alternative: why does it work? Since az = 0, we may proceed vertically
down the left side to compute
pA = ρg(zsurf − zA ) = (1010 kg/m3 ) (9.81 m/s2 ) (0.0214 m + 0.07 m) = 906 Pa
96
Chapter 2 Pressure Distribution in a Fluid
z, k
r, ir
p = pa
Ω
a = –rΩ 2 ir
–a
Still-water
level
Fig. 2.22 Development of paraboloid
constant-pressure surfaces in a fluid
in rigid-body rotation. The dashed
line along the direction of maximum
pressure increase is an exponential
curve.
p = p1
g
g–a
p2
Axis of
rotation
p3
Rigid-Body Rotation
As a second special case, consider rotation of the fluid about the z axis without
any translation, as sketched in Fig. 2.22. We assume that the container has been
rotating long enough at constant Ω for the fluid to have attained rigid-body rotation. The fluid acceleration will then be a centripetal term. In the coordinates of
Fig. 2.22, the angular-velocity and position vectors are given by
Ω = kΩ r0 = irr
(2.41)
Ω × (Ω × r0 ) = −rΩ2ir
(2.42)
Then the acceleration is given by
as marked in the figure, and Eq. (2.38) for the force balance becomes
∇p = ir
∂p
∂p
+k
= ρ(g − a) = ρ(−gk + rΩ2ir )
∂r
∂z
Equating like components, we find the pressure field by solving two first-order
partial differential equations:
∂p
∂p
= ρrΩ2
= −γ
(2.43)
∂r
∂z
The right-hand sides of (2.43) are known functions of r and z. One can proceed as
follows: Integrate the first equation “partially,” holding z constant, with respect to r.
The result is
p = 12ρr2Ω2 + f(z)
(2.44)
2
where the “constant” of integration is actually a function f(z). Now differentiate
this with respect to z and compare with the second relation of (2.43):
∂p
= 0 + f ′(z) = −γ
∂z
2
This is because f(z) vanishes when differentiated with respect to r. If you don’t see this, you
should review your calculus.
2.9 Pressure Distribution in Rigid-Body Motion 97
Still water
level
h
2
Volume =
π
2
R2h
2 2
h= Ω R
2g
h
2
Ω
Fig. 2.23 Determining the freesurface position for rotation of
a cylinder of fluid about its
central axis.
R
R
or
f(z) = −γz + C
where C is a constant. Thus Eq. (2.44) now becomes
p = const − γz + 12ρr2Ω2
(2.45)
This is the pressure distribution in the fluid. The value of C is found by specifying the pressure at one point. If p = p0 at (r, z) = (0, 0), then C = p0. The final
desired distribution is
p = p0 − γz + 12 ρr2Ω2 (2.46)
The pressure is linear in z and parabolic in r. If we wish to plot a constantpressure surface, say, p = p1, Eq. (2.45) becomes
z=
p0 − p1 r2Ω2
= a + br2
+
γ
2g
(2.47)
Thus the surfaces are paraboloids of revolution, concave upward, with their minimum points on the axis of rotation. Some examples are sketched in Fig. 2.22.
As in the previous example of linear acceleration, the position of the free surface
is found by conserving the volume of fluid. For a noncircular container with the
axis of rotation off-center, as in Fig. 2.22, a lot of laborious mensuration is required,
and a single problem will take you all weekend. However, the calculation is easy
for a cylinder rotating about its central axis, as in Fig. 2.23. Since the volume of
a paraboloid is one-half the base area times its height, the still-water level is exactly
halfway between the high and low points of the free surface. The center of the
fluid drops an amount h/2 = Ω2R2/(4g), and the edges rise an equal amount.
EXAMPLE 2.14
The coffee cup in Example 2.13 is removed from the drag racer, placed on a turntable,
and rotated about its central axis until a rigid-body mode occurs. Find (a) the angular
velocity that will cause the coffee to just reach the lip of the cup and (b) the gage
pressure at point A for this condition.
98
Chapter 2 Pressure Distribution in a Fluid
Solution
Part (a)
The cup contains 7 cm of coffee. The remaining distance of 3 cm up to the lip must equal
the distance h/2 in Fig. 2.23. Thus
Solving, we obtain
h
Ω2R2 Ω2 (0.03 m) 2
= 0.03 m =
=
2
4g
4(9.81 m/s2 )
Ω2 = 1308 or
z
Ω = 36.2 rad/s = 345 r/min
Ans. (a)
Part (b)
To compute the pressure, it is convenient to put the origin of coordinates r and z at
the bottom of the free-surface depression, as shown in Fig. E2.14. The gage pressure
here is p0 = 0, and point A is at (r, z) = (3 cm, –4 cm). Equation (2.46) can then be
evaluated:
3 cm
pA = 0 − (1010 kg/m3 ) (9.81 m/s2 ) (−0.04 m)
r
0
+ 12 (1010 kg/m3 ) (0.03 m) 2 (1308 rad2/s2 )
7 cm
= 396 N/m2 + 594 N/m2 = 990 Pa
Ω
Ans. (b)
This is about 43 percent greater than the still-water pressure pA = 694 Pa.
A
3 cm
Here, as in the linear acceleration case, it should be emphasized that the paraboloid pressure distribution (2.46) sets up in any fluid under rigid-body rotation,
regardless of the shape or size of the container. The container may even be closed
and filled with fluid. It is only necessary that the fluid be continuously interconnected throughout the container. The following example will illustrate a peculiar
case in which one can visualize an imaginary free surface extending outside the
walls of the container.
3 cm
E2.14
z
10 in
EXAMPLE 2.15
r
0
30 in
Ω
B
A U-tube with a radius of 10 in and containing mercury to a height of 30 in is rotated
about its center at 180 r/min until a rigid-body mode is achieved. The diameter of the
tubing is negligible. Atmospheric pressure is 2116 lbf/ft2. Find the pressure at point A
in the rotating condition. See Fig. E2.15.
Solution
Convert the angular velocity to radians per second:
A
Ω = (180 r/min)
Imaginary
free surface
E2.15
2π rad/r
= 18.85 rad/s
60 s/min
From Table 2.1 we find for mercury that γ = 846 lbf/ft3 and hence ρ = 846/32.2 =
26.3 slugs/ft3. At this high rotation rate, the free surface will slant upward at a fierce
angle [about 84˚; check this from Eq. (2.47)], but the tubing is so thin that the free
2.9 Pressure Distribution in Rigid-Body Motion 99
surface will remain at approximately the same 30-in height, point B. Placing our origin
of coordinates at this height, we can calculate the constant C in Eq. (2.45) from the
condition pB = 2116 lbf/ft2 at (r, z) = (10 in, 0):
2
2
pB = 2116 lbf/ft2 = C − 0 + 12 (26.3 slugs/ft3 ) ( 10
12 ft) (18.85 rad/s)
C = 2116 − 3245 = −1129 lbf/ft2
or
We then obtain pA by evaluating Eq. (2.46) at (r, z) = (0, –30 in):
2
pA = −1129 − (846 lbf/ft3 ) (−30
12 ft) = −1129 + 2115 = 986 lbf/ft Ans.
This is less than atmospheric pressure, and we can see why if we follow the freesurface paraboloid down from point B along the dashed line in the figure. It will cross
the horizontal portion of the U-tube (where p will be atmospheric) and fall below point
A. From Fig. 2.23 the actual drop from point B will be
h=
2 10 2
Ω2R2 (18.85) ( 12 )
=
= 3.83 ft = 46 in
2g
2(32.2)
2
Thus pA is about 16 inHg below atmospheric pressure, or about 16
12 (846) = 1128 lbf/ft
2
below pa = 2116 lbf/ft , which checks with the answer above. When the tube is at rest,
2
pA = 2116 − 846(−30
12 ) = 4231 lbf/ft
Hence rotation has reduced the pressure at point A by 77 percent. Further rotation can
reduce pA to near-zero pressure, and cavitation can occur.
An interesting by-product of this analysis for rigid-body rotation is that the
lines everywhere parallel to the pressure gradient form a family of curved surfaces, as sketched in Fig. 2.22. They are everywhere orthogonal to the constantpressure surfaces, and hence their slope is the negative inverse of the slope
computed from Eq. (2.47):
dz
dr
∣
GL
=−
1
1
=− 2
(dz/dr) p=const
rΩ /g
where GL stands for gradient line
or
g
dz
=− 2
dr
rΩ
(2.48)
Separating the variables and integrating, we find the equation of the pressuregradient surfaces:
r = C1 exp (−
Ω2z
g )
(2.49)
Notice that this result and Eq. (2.47) are independent of the density of the fluid.
In the absence of friction and Coriolis effects, Eq. (2.49) defines the lines along
which the apparent net gravitational field would act on a particle. Depending on
its density, a small particle or bubble would tend to rise or fall in the fluid along
100
Chapter 2 Pressure Distribution in a Fluid
Fig. 2.24 Experimental
demonstration with buoyant
streamers of the fluid force field
in rigid-body rotation: (top) fluid
at rest (streamers hang vertically
upward); (bottom) rigid-body
rotation (streamers are aligned
with the direction of maximum
pressure gradient). (From ‘The
Apparent Field of Gravity in a
Rotating Fluid System’ by R. Ian
Fletcher. American Journal of
Physics vol. 40, pp. 959–965,
July 1972)
2.10 Pressure Measurement 101
these exponential lines, as demonstrated experimentally in Ref. 5. Also, buoyant
streamers would align themselves with these exponential lines, thus avoiding any
stress other than pure tension. Figure 2.24 shows the configuration of such
streamers before and during rotation.
2.10 Pressure Measurement
Pressure is a derived property. It is the force per unit area as related to fluid
molecular bombardment of a surface. Thus most pressure instruments only
infer the pressure by calibration with a primary device such as a deadweight
piston tester. There are many such instruments, for both a static fluid and a
moving stream. The instrumentation texts in Refs. 6 to 9, 10, 11, and 14–15
list over 20 designs for pressure measurement instruments. These instruments
may be grouped into four categories:
1. Gravity-based: barometer, manometer, deadweight piston.
2. Elastic deformation: bourdon tube (metal and quartz), diaphragm, bellows,
strain-gage, optical beam displacement.
3. Gas behavior: gas compression (McLeod gage), thermal conductance
(Pirani gage), molecular impact (Knudsen gage), ionization, thermal
conductivity, air piston.
4. Electric output: resistance (Bridgman wire gage), diffused strain gage,
capacitative, piezoelectric, potentiometric, magnetic inductance, magnetic reluctance, linear variable differential transformer (LVDT), resonant ­frequency.
5. Luminescent coatings for surface pressures [13].
The gas-behavior gages are mostly special-purpose instruments used for certain
s­cientific experiments. The deadweight tester is the instrument used most often
for calibrations; for example, it is used by the U.S. National Institute for Standards
and Technology (NIST). The barometer is described in Fig. 2.6.
The manometer, analyzed in Sec. 2.4, is a simple and inexpensive hydrostaticprinciple device with no moving parts except the liquid column itself. Manometer measurements must not disturb the flow. The best way to do this is to take
the measurement through a static hole in the wall of the flow, as illustrated in
Fig. 2.25a. The hole should be normal to the wall, and burrs should be avoided.
If the hole is small enough (typically 1-mm diameter), there will be no flow into
the measuring tube once the pressure has adjusted to a steady value. Thus the
flow is almost undisturbed. An oscillating flow pressure, however, can cause a
large error due to possible dynamic response of the tubing. Other devices of
smaller dimensions are used for dynamic-pressure measurements. The manometer
in Fig. 2.25a measures the gage pressure p1. The instrument in Fig. 2.25b is a
digital differential manometer, which can measure the difference between two
different points in the flow, with stated accuracy of 0.3 percent of full scale. The
world of instrumentation is moving quickly toward digital readings.
102
Chapter 2 Pressure Distribution in a Fluid
Flow
p1
Fig. 2.25 Two types of accurate
manometers for precise measurements: (a) tilted tube with eyepiece;
(b) a digital manometer of rated
accuracy ±0.3 percent. (Courtesy
of Sper Scientific)
(a)
(b)
In category 2, elastic-deformation instruments, a popular, inexpensive, and
reliable device is the bourdon tube, sketched in Fig. 2.26. When pressurized
internally, a curved tube with flattened cross section will deflect outward. The
deflection can be measured by a linkage attached to a calibrated dial pointer, as
shown. Or the deflection can be used to drive electric-output sensors, such as a
variable transformer. Similarly, a membrane or diaphragm will deflect under pressure and can either be sensed directly or used to drive another sensor.
An interesting variation of Fig. 2.26 is the fused-quartz, force-balanced bourdon tube, shown in Fig. 2.27, whose spiral-tube deflection is sensed optically and
returned to a zero reference state by a magnetic element whose output is proportional to the fluid pressure. The fused-quartz, force-balanced bourdon tube is
reported to be one of the most accurate pressure sensors ever devised, with uncertainty on the order of ±0.003 percent.
A
Bourdon
tube
Section AA
A
Pointer for
dial gage
Flattened tube deflects
outward under pressure
Linkage
Fig. 2.26 Schematic of a
bourdon-tube device for mechanical
measurement of high pressures.
High pressure
2.10 Pressure Measurement 103
Fluke Calibration Quartz Bourdon Tube (QBT) Pressure Sensor
Fused-quartz bourdon tube
Mirror
Fig. 2.27 The fused-quartz,
force-balanced bourdon tube is the
most accurate pressure sensor used
in commercial applications today.
Electromagnetic coil
Torsion hinge
The quartz gages, both the bourdon type and the resonant type, are expensive
but extremely accurate, stable, and reliable [12]. They are often used for deepocean pressure measurements, which detect long waves and tsunami activity over
extended time periods.
The last category, electric-output sensors, is extremely important in engineering because the data can be stored on computers and freely manipulated, plotted,
and analyzed. Two examples are shown in Fig. 2.28, the first being the capacitive
sensor in Fig. 2.28a. The differential pressure deflects the silicon diaphragm and
changes the capacitance of the liquid in the cavity. Note that the cavity has spherical end caps to prevent overpressure damage. In the second type, a MEMS sensor is arranged to deform under pressure such that its natural vibration frequency
is proportional to the pressure. An oscillator excites the element’s resonant frequency and converts it into appropriate pressure units. Figure 2.28b shows an
autonomous unmanned aerial vehicle (UAV) equipped with integrated 7-axis inertial and barometric pressure sensors for flight control. Figure 2.28c elaborates the
tiny micromachined silicon element.
Another kind of dynamic electric-output sensor is the piezoelectric transducer,
shown in Fig. 2.29. The sensing elements are thin layers of quartz, which generate an electric charge when subjected to stress. The design in Fig. 2.29 is flushmounted on a solid surface and can sense rapidly varying pressures, such as blast
waves. Other designs are of the cavity type. This type of sensor primarily detects
transient pressures, not steady stress, but if highly insulated can also be used for
short-term static events. Note also that it measures gage pressure—that is, it
detects only a change from ambient conditions.
104
Chapter 2 Pressure Distribution in a Fluid
Cover flange
Seal diaphragm
High-pressure side
Low-pressure side
Sensing diaphragm
Filling liquid
(a)
Fig. 2.28 Pressure sensors with
electric output: (a) a silicon
diaphragm whose deflection
changes the cavity capacitance
(b) A drone with MEMS sensors
used in flight controllers and the
camera’s gimbal image stabilizer.
(Courtesy of TDK InvenSense);
(c) a micromachined silicon element that resonates at a frequency
proportional to applied pressure.
Metal
conductors
(100) Si
diaphragm
n-type
epitaxial
layar
(b)
p-type diffused
piezoresistor
Bondpad
R1
R2
R3
(111)
p-type
substrate
and frame
Etched cavity
Backside port
Anodically
bonded
Pyrex
substrate
0 mm
1 mm
2 mm
(c)
3 mm
Problems 105
+
–
Connector
Integrated
circuit
amplifier
Potting
Acceleration
compensation
mass and plate
Fig. 2.29 A piezoelectric transducer
measures rapidly changing pressures.
M
Quartz plates
Source: PCB Piezorronics, Inc, Depew,
New York.
Diaphragm
Summary
This chapter has been devoted entirely to the computation of pressure distributions and the resulting forces and moments in a static fluid or a fluid with a
known velocity field. All hydrostatic (Secs. 2.3 to 2.8) and rigid-body (Sec. 2.9)
problems are solved in this manner and are classic cases that every student should
understand. In arbitrary viscous flows, both pressure and velocity are unknowns
and are solved together as a system of equations in the chapters that follow.
Problems
Most of the problems herein are fairly straightforward. More difficult or open-ended assignments are indicated with an asterisk,
as in Prob. 2.9. Problems labeled with a computer icon
may
require the use of a computer. The standard end-of-chapter problems 2.1 to 2.161 (categorized in the problem distribution) are
followed by word problems W2.1 to W2.9, fundamentals of engineering exam problems FE2.1 to FE2.10, comprehensive problems C2.1 to C2.9, and design projects D2.1 to D2.3.
2.1, 2.2
2.3
2.3
2.4
2.5
Topic
Stresses; pressure gradient; gage pressure
Hydrostatic pressure; barometers
The atmosphere
Manometers; multiple fluids
Forces on plane surfaces
Problems
2.1–2.6
2.7–2.23
2.24–2.29
2.30–2.47
2.48–2.80
Forces on curved surfaces
Forces in layered fluids
Buoyancy; Archimedes’ principles
Stability of floating bodies
Uniform acceleration
Rigid-body rotation
Pressure measurements
2.81–2.100
2.101–2.102
2.103–2.126
2.127–2.136
2.137–2.151
2.152–2.159
2.160–2.161
Stresses; pressure gradient; gage pressure
P2.1
Problem Distribution
Section
2.6
2.7
2.8
2.8
2.9
2.9
2.10
For the two-dimensional stress field shown in Fig. P2.1
it is found that
σxx = 3000 lbf/ft2 σyy = 2000 lbf/ft2 σxy = 500 lbf/ft2
Find the shear and normal stresses (in lbf/ft2) acting on
plane AA cutting through the element at a 30˚ angle as
shown.
106
Chapter 2 Pressure Distribution in a Fluid
P2.6
σyy
σyx
=
σxy
A
σxx
30°
Hydrostatic pressure; barometers
σxx
A
σyx
σyy
P2.1 P2.2
La Paz, Bolivia, is at an altitude of approximately
3658 m. Assume a standard atmosphere. How high
would the liquid rise in a methanol barometer, assumed at 20˚C?
Hint: Don’t forget the vapor pressure.
P2.8 Suppose, which is possible, that there is a half-mile
deep lake of pure ethanol on the surface of Mars.
Estimate the absolute pressure, in Pa, at the bottom
of this speculative lake.
P2.9 A storage tank, 26 ft in diameter and 36 ft high, is filled
with SAE 30W oil at 20˚C. (a) What is the gage pressure, in lbf/in2, at the bottom of the tank? (b) How does
your result in (a) change if the tank diameter is reduced
to 15 ft? (c) Repeat (a) if leakage has caused a layer of 5
ft of water to rest at the bottom of the (full) tank.
P2.10 A large open tank is open to sea-level atmosphere and
filled with liquid, at 20˚C, to a depth of 50 ft. The absolute pressure at the bottom of the tank is approximately
221.5 kPa. From Table A.3, what might this liquid be?
P2.11 In Fig. P2.11, pressure gage A reads 1.5 kPa (gage).
The fluids are at 20˚C. Determine the elevations z, in
meters, of the liquid levels in the open piezometer
tubes B and C.
P2.7
σxy
=
Any pressure reading can be expressed as a length or
head, h = p/ρg. What is standard sea-level pressure expressed in (a) ft of glycerin, (b) inHg, (c) m of water, and
(d ) mm of ethanol? Assume all fluids are at 20˚C.
For the two-dimensional stress field shown in Fig. P2.1
suppose that
σxx = 2000 lbf/ft2 σyy = 3000 lbf/ft2 σn (AA) = 2500 lbf/ft2
Compute (a) the shear stress σxy and (b) the shear stress
on plane AA.
P2.3 A vertical, clean, glass piezometer tube has an inside
diameter of 1 mm. When pressure is applied, water at
20˚C rises into the tube to a height of 25 cm. After
correcting for surface tension, estimate the applied
pressure in Pa.
P2.4 Pressure gages, such as the bourdon gage in Fig. P2.4,
are calibrated with a deadweight piston. If the bourdon
gage is designed to rotate the pointer 10 degrees for
every 2 psig of internal pressure, how many degrees
does the pointer rotate if the piston and weight together
total 44 newtons?
A
W
2 cm
diameter
θ?
B
Bourdon
gage
Oil
2m
Air
1.5 m
Gasoline
1m
Glycerin
P2.4 P2.5
Quito, Ecuador, has an average altitude of 2850 m. On a
standard day, pressure gage A in a laboratory experiment
reads 63 kPa and gage B reads 105 kPa. Express these readings in gage pressure or vacuum pressure, whichever is
appropriate.
P2.11 C
z=0
P2.12 In Fig. P2.12 the tank contains water and immiscible
oil at 20˚C. What is h in cm if the density of the oil is
898 kg/m3?
Problems 107
15 lbf/in2 abs
h
A
6 cm
12 cm
Air
2 ft
Oil
8 cm
1 ft
Oil
Water
1 ft
P2.12 P2.13 In Fig. P2.13 the 20˚C water and gasoline surfaces are
open to the atmosphere and at the same elevation. What
is the height h of the third liquid in the right leg?
1.5 m
Water
2 ft
C
P2.15 P2.16 If the absolute pressure at the interface between water and
mercury in Fig. P2.16 is 93 kPa, what, in lbf/ft2, is (a) the
pressure at the surface and (b) the pressure at the bottom of
the container?
Gasoline
Water
B
h
1m
P2.13
P2.14 For the three-liquid system shown, compute h1 and h2.
­Neglect the air density.
Water
h2
Oil,
SG =
0.78
27 cm
8 cm
h1
75°
75°
Mercury
28 cm
Water
Liquid, SG = 1.60
32 cm
P2.16
P2.17 The system in Fig. P2.17 is at 20˚C. Determine the
height h of the water in the left side.
0 Pa (gage)
5 cm
Air, 200 Pa (gage)
25 cm
P2.14
P2.15 The air–oil–water system in Fig. P2.15 is at 20˚C. Knowing that gage A reads 15 lbf/in2 absolute and gage B reads
1.25 lbf/in2 less than gage C, compute (a) the specific
weight of the oil in lbf/ft3 and (b) the actual reading of
gage C in lbf/in2 absolute.
8 cm
Mercury
h
P2.17
Oil, SG = 0.8
Water
20 cm
108
Chapter 2 Pressure Distribution in a Fluid
P2.18 The system in Fig. P2.18 is at 20˚C. If atmospheric
pressure is 101.33 kPa and the pressure at the bottom
of the tank is 242 kPa, what is the specific gravity of
fluid X?
2000
lbf
3-in diameter
1 in
SAE 30 oil
Water
F
1-in diameter
1m
2m
15 in
Oil
P2.20
Fluid X
3m
Air: 180 kPa abs
Mercury
0.5 m
P2.19 The U-tube in Fig. P2.19 has a 1-cm ID and contains
mercury as shown. If 20 cm3 of water is poured into the
right-hand leg, what will the free-surface height in each
leg be after the sloshing has died down?
Water
h?
P2.18 80 cm
Mercury
A
P2.21
B
P2.22 The fuel gage for a gasoline tank in a car reads proportional to the bottom gage pressure as in Fig.
P2.22. If the tank is 30 cm deep and accidentally
contains 2 cm of water plus gasoline, how many centimeters of air remain at the top when the gage erroneously reads “full”?
Mercury
10 cm
10 cm
Vent
Air
P2.19 10 cm
P2.20 The hydraulic jack in Fig. P2.20 is filled with oil at 56 lbf/
ft3. Neglecting the weight of the two pistons, what force
F on the handle is required to support the 2000-lbf
weight for this design?
P2.21 At 20˚C gage A reads 350 kPa absolute. What is the
height h of the water in cm? What should gage B read in
kPa ­absolute? See Fig. P2.21.
30 cm
Gasoline
SG = 0.68
Water
P2.22 h?
2 cm
pgage
P2.23 In Fig. P2.23 both fluids are at 20˚C. If surface tension effects are negligible, what is the density of the
oil, in kg/m3?
Problems 109
Oil
6 cm
8 cm
inverted, so the original top rim of the glass is at the bottom
of the picture, and the original bottom of the glass is at the
top of the picture. The weight of the card can be neglected.
(c) Estimate the theoretical maximum glass height at which
this experiment could still work, such that the water would
not fall out of the glass.
Card
Water
Top of glass
10 cm
P2.23 The atmosphere
P2.24 In Prob. 1.2 we made a crude integration of the density
distribution ρ(z) in Table A.6 and estimated the mass
of the earth’s atmosphere to be m ≈ 6 E18 kg. Can this
­result be used to estimate sea-level pressure on the
earth? Conversely, can the actual sea-level pressure of
101.35 kPa be used to make a more accurate estimate
of the atmospheric mass?
*P2.25 As measured by NASA’s Viking landers, the atmosphere of Mars, where g ≈3.71 m/s2, is almost entirely
carbon dioxide, and the surface pressure averages 700 Pa.
The temperature is cold and drops off exponentially:
T ≈ To e−Cz, where C = 1.3E-5 m−1 and To = 250 K. For
example, at 20,000 m altitude, T ≈ 193 K. (a) Find an
­analytic formula for the variation of pressure with
altitude. (b) Find the altitude where pressure on Mars
has dropped to 1 pascal.
P2.26 For gases that undergo large changes in height, the linear
approximation, Eq. (2.14), is inaccurate. Expand the
troposphere power-law, Eq. (2.20), into a power series,
and show that the linear approximation p ≈ pa – ρa gz is
adequate when
δz ≪
2T0
(n − 1)B
where n =
g
RB
P2.27 Conduct an experiment to illustrate atmospheric pressure. Note: Do this over a sink or you may get wet! Find
a ­drinking glass with a very smooth, uniform rim at the
top. Fill the glass nearly full with water. Place a smooth,
light, flat plate on top of the glass such that the entire
rim of the glass is covered. A glossy postcard works
best. A small index card or one flap of a greeting card
will also work. See Fig. P2.27a.
(a) Hold the card against the rim of the glass and turn the
glass upside down. Slowly release pressure on the card.
Does the water fall out of the glass? Record your experimental ­observations. (b) Find an expression for the pressure
at points 1 and 2 in Fig. P2.27b. Note that the glass is now
Bottom of glass
P2.27a Original bottom of glass
1●
2●
P2.27b Card
Original top of glass
P2.28 A correlation of computational fluid dynamics results
indicates that, all other things being equal, the distance
traveled by a well-hit baseball varies inversely as the
0.36 power of the air density: that is, X/X0 = (ρ0/ρ)0.36,
where X0 and ρ0 are the distance and density at standard
conditions, X and ρ are the local distance and density. If
a home-run ball hit in Citi Field in New York travels 400
ft, estimate the distance it would travel in (a) Quito, Ecuador, and (b) Colorado Springs, CO.
P2.29 Follow up on Prob. P2.8 by estimating the altitude on
Mars where the pressure has dropped to 20 percent of its
surface value. Assume an isothermal atmosphere, not
the exponential variation of P2.25.
Manometers; multiple fluids
P2.30 For the traditional equal-level manometer measurement
in Fig. E2.3, water at 20˚C flows through the plug device
from a to b. The manometer fluid is mercury. If L = 12
cm and h = 24 cm, (a) what is the pressure drop through
the device? (b) If the water flows through the pipe at a
velocity V = 18 ft/s, what is the dimensionless loss coefficient of the device, defined by K = Δp/(ρV2)? We will
study loss coefficients in Chap. 6.
110
Chapter 2 Pressure Distribution in a Fluid
P2.31 In Fig. P2.31 all fluids are at 20˚C. Determine the pres- * P2.34 Sometimes manometer dimensions have a significant
sure difference (Pa) between points A and B.
­effect. In Fig. P2.34 containers (a) and (b) are cylindrical
and conditions are such that pa = pb. Derive a formula for
the pressure difference pa – pb when the oil–water interKerosene
face on the right rises a distance Δh < h, for (a) d ≪ D
Air
and (b) d = 0.15D. What is the percentage change in the
Benzene
value of Δp?
B
40 cm
A
20 cm
Mercury
9 cm
D
14 cm
8 cm
D
(b)
Water
(a)
P2.31
P2.32 For the inverted manometer of Fig. P2.32, all fluids
are at 20˚C. If pB – pA = 97 kPa, what must the height
H be in cm?
L
SAE 30 oil
H
Water
h
Meriam
red oil,
SG = 0.827
18 cm
Water
d
H
Mercury
A
P2.34
P2.35 Water flows upward in a pipe slanted at 30˚, as in
Fig. P2.35. The mercury manometer reads h = 12 cm.
Both fluids are at 20˚C. What is the pressure difference
p1 – p2 in the pipe?
35 cm
B
P2.32 (2)
2
P2.33 In Fig. P2.33 the pressure at point A is 25 lbf/in . All
fluids are at 20˚C. What is the air pressure in the closed
chamber B, in Pa?
30°
(1)
h
Air B
SAE 30 oil
3 cm
A
Water
4 cm
6 cm
8 cm
3 cm
P2.33
Liquid, SG = 1.45 5 cm
P2.35 2m
P2.36 In Fig. P2.36 both the tank and the tube are open to the
­atmosphere. If L = 2.13 m, what is the angle of tilt θ of
the tube?
P2.37 The inclined manometer in Fig. P2.37 contains Meriam
red manometer oil, SG = 0.827. Assume that the reservoir is very large. If the inclined arm is fitted with graduations 1 in apart, what should the angle θ be if each
graduation corresponds to 1 lbf/ft2 gage pressure for pA?
Problems 111
Air
50 cm
Oil
SG = 0.8
50 cm
Water
SG = 1.0
8 cm
8 cm
L
12 cm
Oil,
SG = 0.8
θ
9 cm
P2.36
11 cm
Mercury
P2.39
1 in
pA
D=
θ
5
16
B
in
3 ft
1 ft
Air
2 ft
Reservoir
P2.37
4 ft
P2.38 If the pressure in container A in Fig. P2.38 is 200 kPa,
compute the pressure in container B.
Water
A
P2.40 Oil
P2.41 The system in Fig. P2.41 is at 20˚C. Compute the pressure at point A in lbf/ft2 absolute.
B
Water
Water
18 cm
Oil, SG = 0.85
A
16 cm
A
Oil,
SG = 0.8
Mercury
5 in
6 in
8 cm
P2.39 In Fig. P2.39 the right leg of the manometer is open to
the atmosphere. Find the gage pressure, in Pa, in the air
gap in the tank.
P2.40 In Fig. P2.40, if pressure gage A reads 20 lbf/in2 absolute, find the pressure in the closed air space B. The manometer fluid is Meriam red oil, SG = 0.827.
10 in
Water
22 cm
P2.38 pa = 14.7 lbf/in2
Mercury
P2.41
P2.42 Very small pressure differences pA – pB can be measured accurately by the two-fluid differential manometer in Fig. P2.42. Density ρ2 is only slightly larger
than that of the upper fluid ρ1. Derive an expression
for the proportionality between h and pA – pB if the
reservoirs are very large.
112
Chapter 2 Pressure Distribution in a Fluid
pA
pB
ρ1
ρ1
P2.45 In Fig. P2.45, determine the gage pressure at point A in
Pa. Is it higher or lower than atmospheric?
patm
Air
h1
h1
h
Oil,
SG = 0.85
ρ
2
30 cm
P2.42 45 cm
P2.43 The traditional method of measuring blood pressure uses
a sphygmomanometer, first recording the highest (systolic) and then the lowest (diastolic) pressure from which
flowing “Korotkoff” sounds can be heard. P
­ atients with
dangerous hypertension can exhibit systolic ­pressures as
high as 5 lbf/in2. Normal levels, however, are 2.7 and
1.7 lbf/in2, respectively, for systolic and d­ iastolic pressures. The ­manometer uses mercury and air as fluids.
(a) How high in cm should the manometer tube be?
(b) Express normal systolic and diastolic blood pressure
in millimeters of mercury.
P2.44 Water flows downward in a pipe at 45˚, as shown in
Fig. P2.44. The pressure drop p1 – p2 is partly due to
gravity and partly due to friction. The mercury
manometer reads a 6-in height difference. What is the
total pressure drop p1 – p2 in lbf/in2? What is the
pressure drop due to friction only between 1 and 2 in
lbf/in2? Does the manometer reading correspond only
to friction drop? Why?
40 cm
15 cm
A
Water
P2.45
Mercury
P2.46 In Fig. P2.46 both ends of the manometer are open to the
atmosphere. Estimate the specific gravity of fluid X.
10 cm
SAE 30 oil
9 cm
Water
5 cm
7 cm
45°
1
5 ft
4 cm
Fluid X
6 cm
Flow
2
Water
6 in
Mercury
P2.44 P2.46 12 cm
P2.47 The cylindrical tank in Fig. P2.47 is being filled with water at
20˚C by a pump developing an exit pressure of 175 kPa. At
the instant shown, the air pressure is 110 kPa and H = 35 cm.
The pump stops when it can no longer raise the water pressure. For isothermal air compression, estimate H at that time.
Problems 113
50 cm
Newspaper
75 cm
Air
20° C
Ruler
H
P2.47 Desk
Water
Pump
P2.48 The system in Fig. P2.48 is open to 1 atm on the right
side. (a) If L = 120 cm, what is the air pressure in container A? (b) Conversely, if pA = 135 kPa, what is the
length L?
Air
A
32 cm
Mercury
P2.50 A small submarine, with a hatch door 30 in in diameter, is submerged in seawater. (a) If the water hydrostatic force on the hatch is 69,000 lbf, how deep is the
sub? (b) If the sub is 350 ft deep, what is the hydrostatic force on the hatch?
P2.51 Gate AB in Fig. P2.51 is 1.2 m long and 0.8 m into the
­paper. Neglecting atmospheric pressure, compute the
force F on the gate and its center-of-pressure p­ osition X.
6m
L
18 cm
15 cm
P2.49
35˚
Oil,
SG = 0.82
4m
Water
P2.48
8m
1m
Forces on plane surfaces
P2.49 Conduct the following experiment to illustrate air pressure. Find a thin wooden ruler (approximately 1 ft in
length) or a thin wooden paint stirrer. Place it on the
edge of a desk or table with a little less than half of it
hanging over the edge lengthwise. Get two full-size
sheets of newspaper; open them up and place them on
top of the ruler, covering only the portion of the ruler
resting on the desk as illustrated in Fig. P2.49. (a) Estimate the total force on top of the newspaper due to air
pressure in the room. (b) Careful! To avoid potential
injury, make sure nobody is standing directly in front of
the desk. Perform a karate chop on the portion of the
ruler sticking out over the edge of the desk. Record
your results. (c) Explain your results.
A
X
1.2 m
B
F
P2.51
40°
P2.52 Example 2.5 calculated the force on plate AB and its
line of action, using the moment-of-inertia approach.
Some teachers say it is more instructive to calculate
these by direct integration of the pressure forces. Using Figs. P2.52 and E2.5a, (a) find an expression for
the pressure variation p(ξ ) along the plate; (b) integrate this ­expression to find the total force F; (c) integrate the ­moments about point A to find the position
of the center of pressure.
114
Chapter 2 Pressure Distribution in a Fluid
ξ
p(ξ)
P2.57 The square vertical panel ABCD in Fig. P2.57 is submerged in water at 20˚C. Side AB is at least 1.7 m below
the ­surface. Determine the difference between the hydrostatic forces on subpanels ABD and BCD.
A
6 ft
B
P2.52
A
B
8 ft
P2.53 The Hoover Dam, in Arizona, encloses Lake Mead,
which contains 10 trillion gallons of water. The dam is
1200 ft wide and the lake is 500 ft deep. (a) Estimate
the hydrostatic force on the dam, in MN. (b) Explain
how you might analyze the stress in the dam due to this
hydrostatic force.
P2.54 In Fig. P2.54, the hydrostatic force F is the same on the
bottom of all three containers, even though the weights
of liquid above are quite different. The three bottom
shapes and the fluids are the same. This is called the
hydrostatic paradox. Explain why it is true and sketch a
free body of each of the liquid columns.
60 cm
P2.57
D
C
P2.58 In Fig. P2.58, the cover gate AB closes a circular opening
80 cm in diameter. The gate is held closed by a 200-kg
mass as shown. Assume standard gravity at 20˚C. At
what water level h will the gate be dislodged? Neglect
the weight of the gate.
200 kg
h
m
B
F
P2.54 F
(a)
F
(b)
30 cm
A
Water
3m
P2.58
(c)
*P2.59 Gate AB has length L and width b into the paper, is
hinged at B, and has negligible weight. The liquid level h
P2.55 Gate AB in Fig. P2.55 is 5 ft wide into the paper, hinged
remains at the top of the gate for any angle θ. Find an
at A, and restrained by a stop at B. The water is at 20˚C.
analytic ­expression for the force P, perpendicular to AB,
­Compute (a) the force on stop B and (b) the reactions at
required to keep the gate in equilibrium in Fig. P2.59.
A if the water depth h = 9.5 ft.
P
A
pa
Water
h
pa
L
h
Hinge
A
4 ft
B
P2.55
P2.56 In Fig. P2.55, gate AB is 5 ft wide into the paper, and
stop B will break if the water force on it equals 9200 lbf.
For what water depth h is this condition reached?
P2.59 θ
B
P2.60 In Fig. P2.60, vertical, unsymmetrical trapezoidal panel
ABCD is submerged in fresh water with side AB 12 ft
­below the surface. Since trapezoid centroid formulas are
complicated, (a) estimate, reasonably, the water force on
the panel, in lbf, neglecting atmospheric pressure. For
extra credit, (b) look up the centroid formula and compute the exact force on the panel.
Problems 115
6 ft
A
B
Water
8 ft
C
H
D
9 ft
P2.60 50°
*P2.61 Gate AB in Fig. P2.61 is a homogeneous mass of 180 kg,
1.2 m wide into the paper, hinged at A, and resting on a
smooth bottom at B. All fluids are at 20˚C. For what
water depth h will the force at point B be zero?
h
2 cm
Plug,
D = 4 cm
Mercury
P2.63
*P2.64 Gate ABC in Fig. P2.64 has a fixed hinge line at B and is
2 m wide into the paper. The gate will open at A to release water if the water depth is high enough. Compute
the depth h for which the gate will begin to open.
Water
Glycerin
h
2m
C
A
1m
B
A
60°
P2.61
P2.62 Gate AB in Fig. P2.62 is 15 ft long and 8 ft wide into the
paper and is hinged at B with a stop at A. The water is at
20˚C. The gate is 1-in-thick steel, SG = 7.85. Compute
the water level h for which the gate will start to fall.
Pulley
A
1m
Water at 20°C
P2.64
Water
15 ft
P2.62
h
*P2.65 Gate AB in Fig. P2.65 is semicircular, hinged at B, and
held by a horizontal force P at A. What force P is required for equilibrium?
10,000 lb
60°
B
20 cm
h
5m
B
P2.63 The tank in Fig. P2.63 has a 4-cm-diameter plug at the
­bottom on the right. All fluids are at 20˚C. The plug will
pop out if the hydrostatic force on it is 25 N. For this condition, what will be the reading h on the mercury manometer
on the left side?
Water
A
3m
B
P2.65 P
Gate:
Side view
116
Chapter 2 Pressure Distribution in a Fluid
P2.66 Dam ABC in Fig. P2.66 is 30 m wide into the paper and
made of concrete (SG = 2.4). Find the hydrostatic force
on surface AB and its moment about C. Assuming no
seepage of water under the dam, could this force tip the
dam over? How does your argument change if there is
seepage under the dam?
A
80 m
A
θ
B
F
Water specific weight γ
P2.69
P2.70 The swing-check valve in Fig. P2.70 covers a 22.86-cm
diameter opening in the slanted wall. The hinge is 15 cm
from the centerline, as shown. The valve will open when
the hinge moment is 50 N · m. Find the value of h for the
water to cause this condition.
Water 20°C
Dam
Air
B
C
15 cm
60 m
P2.66
h
Hinge
*P2.67 Generalize Prob. P2.66 as follows. Denote length AB as
Water at 20°C
H, length BC as L, and angle ABC as θ. Let the dam mate60°
P2.70 rial have specific gravity SG. The width of the dam is b.
Assume no seepage of water under the dam. Find an ana- *P2.71 In Fig. P2.71 gate AB is 3 m wide into the paper and is
lytic relation between SG and the critical angle θc for which
c­ onnected by a rod and pulley to a concrete sphere (SG
the dam will just tip over to the right. Use your relation to
= 2.40). What diameter of the sphere is just sufficient to
compute θc for the special case SG = 2.4 (concrete).
keep the gate closed?
P2.68 Isosceles triangle gate AB in Fig. P2.68 is hinged at A
Concrete
and weighs 1500 N. What horizontal force P is required
sphere, SG = 2.4
at point B for equilibrium?
6m
A
Oil, SG = 0.83
3m
8m
1m
4m
A
Water
B
Gate
P2.71
2m
50° B
P2.72 In Fig. P2.72, gate AB is circular. Find the moment of the
hydrostatic force on this gate about axis A.
P
P2.68
P2.69 Consider the slanted plate AB of length L in Fig. P2.69.
(a) Is the hydrostatic force F on the plate equal to the
weight of the missing water above the plate? If not, correct
this ­hypothesis. Neglect the atmosphere. (b) Can a “missing
­water” theory be generalized to curved surfaces of this type?
3m
Water
A
B
P2.72
2m
Problems 117
P2.73 Gate AB is 5 ft wide into the paper and opens to let fresh
water out when the ocean tide is dropping. The hinge at
A is 2 ft above the freshwater level. At what ocean level
h will the gate first open? Neglect the gate weight.
50°
A
Tide
range
3m
C
3m
10 ft
h
Seawater, SG = 1.025
P2.73
3m
B
Water
at 20°C
Stop
P2.76
B
pa
P2.74 Find the height H in Fig. P2.74 for which the hydrostatic
force on the rectangular panel is the same as the force on
the semicircular panel below.
Water
pa
h
A
H
1m
2R
B
1m
P2.74 C
P2.75 The cap at point B on the 5-cm-diameter tube in Fig. P2.75
will be dislodged when the hydrostatic force on its base
reaches 22 lbf. For what water depth h does this occur?
P
P2.77
Oil,
SG = 0.8
B
1m
Water
h
P2.76 Panel BC in Fig. P2.76 is circular. Compute (a) the hydrostatic force of the water on the panel, (b) its center of
pressure, and (c) the moment of this force about point B.
P2.77 The circular gate ABC in Fig. P2.77 has a 1-m radius and
is hinged at B. Compute the force P just sufficient to
keep the gate from opening when h = 8 m. Neglect atmospheric pressure.
P2.78 Panels AB and CD in Fig. P2.78 are each 120 cm wide
into the paper. (a) Can you deduce, by inspection, which
panel has the larger water force? (b) Even if your deduction is brilliant, calculate the panel forces anyway.
D
A
2m
P2.75 30 cm
Water
40 cm
50 cm
40 cm
40°
B
C
50°
P2.78
P2.79 Gate ABC in Fig. P2.79 is 1 m square and is hinged at B.
It will open automatically when the water level h becomes high enough. Determine the lowest height for
which the gate will open. Neglect atmospheric pressure.
Is this result independent of the liquid density?
118
Chapter 2 Pressure Distribution in a Fluid
F
B
A
Water
h
A
60 cm
C
40 cm
Water
r = 8 ft
P2.83 P2.79
*P2.80 A concrete dam (SG = 2.5) is made in the shape of an
isosceles triangle, as in Fig. P2.80. Analyze this geometry to find the range of angles θ for which the hydrostatic
force will tend to tip the dam over at point B. The width
into the paper is b.
B
P2.84 Panel AB in Fig. P2.84 is a parabola with its maximum
at point A. It is 150 cm wide into the paper. Neglect atmospheric pressure. Find (a) the vertical and (b) the
horizontal water forces on the panel.
25 cm
Water
A
Parabola
75 cm
h
40 cm
P2.84 P2.85 Compute the horizontal and vertical components of the
­hydrostatic force on the quarter-circle panel at the bottom of the water tank in Fig. P2.85.
θ B
P2.80 θ
B
6m
Forces on curved surfaces
P2.81 For the semicircular cylinder CDE in Example 2.9, find
the vertical hydrostatic force by integrating the vertical
component of pressure around the surface from θ = 0 to
θ = π.
*P2.82 The dam in Fig. P2.82 is a quarter circle 50 m wide into
the paper. Determine the horizontal and vertical components of the hydrostatic force against the dam and the
point CP where the resultant strikes the dam.
20 m
20 m
pa = 0
CP
5m
Water
2m
P2.85
P2.86 The quarter circle gate BC in Fig. P2.86 is hinged at C.
Find the horizontal force P required to hold the gate
­stationary. Neglect the weight of the gate.
P
B
P2.82 Water
*P2.83 Gate AB in Fig. P2.83 is a quarter circle 10 ft wide into
the paper and hinged at B. Find the force F just sufficient
to keep the gate from opening. The gate is uniform and
weighs 3000 lbf.
2m
2m
P2.86
C
Water
Problems 119
P2.87 The bottle of champagne (SG = 0.96) in Fig. P2.87 is
­under pressure, as shown by the mercury-manometer reading. Compute the net force on the 2-in-radius hemispherical end cap at the bottom of the bottle.
60 cm
p = 200 kPa
30 cm
B
Benzene
at 20°C
60 cm
A
P2.89
4 in
2 in
6 in
150 cm
r = 2 in
P2.87
Mercury
A
*P2.88 Gate ABC is a circular arc, sometimes called a Tainter
gate, which can be raised and lowered by pivoting about
point O. See Fig. P2.88. For the position shown, determine (a) the hydrostatic force of the water on the gate and
(b) its line of action. Does the force pass through point O?
6m
R=6m
B
B
40 cm
P2.90
P2.91 The hemispherical dome in Fig. P2.91 weighs 30 kN and
is filled with water and attached to the floor by six
equally spaced bolts. What is the force in each bolt required to hold down the dome?
C
Water
75 cm
3 cm
O
4m
6m
A
Six
bolts
P2.88
P2.89 The tank in Fig. P2.89 contains benzene and is pressurized to 200 kPa (gage) in the air gap. Determine the vertical hydrostatic force on circular-arc section AB and its
line of action.
P2.90 The tank in Fig. P2.90 is 120 cm long into the paper.
Determine the horizontal and vertical hydrostatic forces
on the quarter-circle panel AB. The fluid is water at
20˚C. Neglect atmospheric pressure.
Water
2m
P2.91
P2.92 A 4-m-diameter water tank consists of two half cylinders, each weighing 4.5 kN/m, bolted together as shown
in Fig. P2.92. If the support of the end caps is neglected,
­determine the force induced in each bolt.
120
Chapter 2 Pressure Distribution in a Fluid
2m
A
h
R
Water
Bolt spacing 25 cm
R
2m
Water
P2.92 *P2.93 In Fig. P2.93, a one-quadrant spherical shell of radius R
is submerged in liquid of specific weight γ and depth
h > R. Find an analytic expression for the resultant hydrostatic force, and its line of action, on the shell surface.
z
O
P2.95 P2.96 In Fig. P2.96, curved section AB is 5 m wide into the
paper and is a 60˚ circular arc of radius 2 m. Neglecting
atmospheric pressure, calculate the vertical and horizontal ­hydrostatic forces on arc AB.
ρ, γ
h
R
Water
4m
B
R
y
A
R
C
P2.93 x
P2.94 Find an analytic formula for the vertical and horizontal
forces on each of the semicircular panels AB in Fig.
P2.94. The width into the paper is b. Which force is
larger? Why?
h
P2.97 The contractor ran out of gunite mixture and finished the
deep corner of a 5-m-wide swimming pool with a quarter-circle piece of PVC pipe, labeled AB in Fig. P2.97.
­Compute the horizontal and vertical water forces on the
curved panel AB.
h
A
A
ρ
d
60°
O
P2.96
+
B
d
+
B
P2.94
*P2.95 The uniform body A in Fig. P2.95 has width b into the
­paper and is in static equilibrium when pivoted about
hinge O. What is the specific gravity of this body if (a) h
= 0 and (b) h = R?
Water
2m
ρ
A
P2.97
1m
B
P2.98 The curved surface in Fig. P2.98 consists of two quarterspheres and a half cylinder. A side view and front view
are shown. Calculate the horizontal and vertical forces
on the surface.
Problems 121
Water
1.5 m
Side
2m
30 cm
Front
1m
80 cm
P2.98
P2.99 The mega-magnum cylinder in Fig. P2.99 has a hemispherical bottom and is pressurized with air to 75 kPa
(gage) at the top. Determine (a) the horizontal and (b)
the vertical hydrostatic forces on the hemisphere, in lbf.
Water
20 ft
12 ft
P2.100 Pressurized water fills the tank in Fig. P2.100. Compute
the net hydrostatic force on the conical surface ABC.
SAE 30W oil
A
Water
90 cm
P2.101
Air
P2.99
Air
60 cm
1m
C
B
160 cm
P2.102 A cubical tank is 3 m × 3 m × 3 m and is layered with
1 meter of fluid of specific gravity 1.0, 1 meter of fluid
with SG = 0.9, and 1 meter of fluid with SG = 0.8. Neglect ­atmospheric ­pressure. Find (a) the hydrostatic
force on the ­bottom and (b) the force on a side panel.
Buoyancy; Archimedes’ principles
P2.103 A solid block, of specific gravity 0.9, floats such that
75 percent of its volume is in water and 25 percent of its
volume is in fluid X, which is layered above the water.
What is the specific gravity of fluid X?
P2.104 The can in Fig. P2.104 floats in the position shown.
What is its weight in N?
3 cm
2m
A
C
8 cm
Water
4m
B
P2.100
7m
150 kPa
gage
Water
Forces on layered surfaces
P2.101 The closed layered box in Fig. P2.101 has square horizontal cross sections everywhere. All fluids are at 20˚C.
­Estimate the gage pressure of the air if (a) the hydrostatic
force on panel AB is 48 kN or (b) the hydrostatic force on
the bottom panel BC is 97 kN.
P2.104
D = 9 cm
P2.105 It is said that Archimedes discovered the buoyancy laws
when asked by King Hiero of Syracuse to determine
whether his new crown was pure gold (SG = 19.3).
­Archimedes measured the weight of the crown in air to be
11.8 N and its weight in water to be 10.9 N. Was it pure gold?
P2.106 A spherical helium balloon has a total mass of 3 kg. It
settles in a calm standard atmosphere at an altitude of
5500 m. Estimate the diameter of the balloon.
P2.107 Repeat Prob. 2.62, assuming that the 10,000-lbf weight
is aluminum (SG = 2.71) and is hanging submerged in
the water.
122
Chapter 2 Pressure Distribution in a Fluid
P2.108 A 7-cm-diameter solid aluminum ball (SG = 2.7) and a
solid brass ball (SG = 8.5) balance nicely when submerged in a liquid, as in Fig. P2.108. (a) If the fluid is
water at 20˚C, what is the diameter of the brass ball? (b)
If the brass ball has a diameter of 3.8 cm, what is the
density of the fluid?
1m
D = 8 cm
θ
Water at 20°C
2 pulleys
4m
+
+
String
P2.108
Brass
Aluminum
D = 7 cm
P2.109 A hydrometer floats at a level that is a measure of the
­specific gravity of the liquid. The stem is of constant diameter D, and a weight in the bottom stabilizes the body
to float vertically, as shown in Fig. P2.109. If the position h = 0 is pure water (SG = 1.0), derive a formula for
h as a function of total weight W, D, SG, and the specific
weight γ0 of water.
P2.112
P2.113 A spar buoy is a buoyant rod weighted to float and protrude vertically, as in Fig. P2.113. It can be used for measurements or markers. Suppose that the buoy is maple
wood (SG = 0.6), 2 in by 2 in by 10 ft, floating in seawater (SG = 1.025). How many pounds of steel (SG = 7.85)
should be added to the bottom end so that h = 18 in?
h
D
SG = 1.0
h
Wsteel
Fluid, SG > 1
W
P2.109 P2.110 A solid sphere, of diameter 18 cm, floats in 20˚C water
with 1527 cubic centimeters exposed above the surface.
(a) What are the weight and specific gravity of this
sphere? (b) Will it float in 20˚C gasoline? If so, how
many cubic centimeters will be exposed?
P2.111 A solid wooden cone (SG = 0.729) floats in water. The
cone is 30 cm high, its vertex angle is 90˚, and it floats
with vertex down. How much of the cone protrudes
above the water?
P2.112 The uniform 5-m-long round wooden rod in Fig. P2.112
is tied to the bottom by a string. Determine (a) the tension in the string and (b) the specific gravity of the wood.
Is it possible for the given information to determine the
inclination angle θ? Explain.
P2.113 P2.114 The uniform rod in Fig. P2.114 is hinged at point B on
the waterline and is in static equilibrium as shown when
2 kg of lead (SG = 11.4) is attached to its end. What is
the specific gravity of the rod material? What is peculiar
about the rest angle θ = 30˚?
Hinge
B
D = 4 cm
θ = 30°
8m
2 kg of lead
P2.114
Problems 123
P2.115 The 2-in by 2-in by 12-ft spar buoy from Fig. P2.113 has
5 lbm of steel attached and has gone aground on a rock, as
in Fig. P2.115. Compute the angle θ at which the buoy will
lean, assuming that the rock exerts no moments on the spar.
5 lbf
θ
9 ft
Water
4 in × 4 in
P2.119
8 ft
θ
Wood
SG = 0.6
P2.120 A uniform wooden beam (SG = 0.65) is 10 cm by 10 cm
by 3 m and is hinged at A, as in Fig. P2.120. At what
angle θ will the beam float in the 20˚C water?
A
Seawater
A
1m
Rock
θ
P2.115
P2.116 The bathysphere of the chapter-opener photo is steel,
SG ≈ 7.85, with inside diameter 54 inches and wall
thickness 1.5 inches. Will the empty sphere float in seawater?
P2.117 The solid sphere in Fig. P2.117 is iron (SG ≈ 7.9). The
tension in the cable is 600 lbf. Estimate the diameter of
the sphere, in cm.
Water
P2.120
P2.121 The uniform beam in Fig. P2.121, of size L by h by b and
with specific weight γb, floats exactly on its diagonal when
a heavy uniform sphere is tied to the left corner, as shown.
Width b << L
L
h << L
γb
γ
Water
SG > 1
P2.117
P2.118 An intrepid treasure-salvage group has discovered a steel
box, containing gold doubloons and other valuables,
resting in 80 ft of seawater. They estimate the weight of
the box and treasure (in air) at 7000 lbf. Their plan is to
attach the box to a sturdy balloon, inflated with air to 3
atm pressure. The empty balloon weighs 250 lbf. The
box is 2 ft wide, 5 ft long, and 18 in high. What is the
proper diameter of the balloon to ensure an upward lift
force on the box that is 20 percent more than required?
P2.119 When a 5-lbf weight is placed on the end of the uniform
floating wooden beam in Fig. P2.119, the beam tilts at an
angle θ with its upper right corner at the surface, as
shown. Determine (a) the angle θ and (b) the specific
gravity of the wood. Hint: Both the vertical forces and
the moments about the beam centroid must be balanced.
Diameter D
P2.121
Show that this can happen only (a) when γb = γ/3 and
(b) when the sphere has size
D =[
1/3
Lhb
]
π(SG − 1)
P2.122 A uniform block of steel (SG = 7.85) will “float” at a
­mercury–water interface as in Fig. P2.122. What is the
­ratio of the distances a and b for this condition?
Water
Steel
block
P2.122 Mercury: SG = 13.56
a
b
124
Chapter 2 Pressure Distribution in a Fluid
P2.123 A barge has the trapezoidal shape shown in Fig. P2.123
and is 22 m long into the paper. If the total weight of
barge and cargo is 350 tons, what is the draft H of the
barge when floating in seawater?
H
60°
60°
h
2.5 m
8m
P2.123 P2.124 A balloon weighing 3.5 lbf is 6 ft in diameter. It is filled
with hydrogen at 18 lbf/in2 absolute and 60˚F and is
­released. At what altitude in the U.S. standard atmosphere will this balloon be neutrally buoyant?
P2.125 A uniform cylindrical white oak log, ρ = 710 kg/m3, floats
lengthwise in fresh water at 20˚C. Its diameter is 24
inches. What height of the log is visible above the surface?
P2.126 A block of wood (SG = 0.6) floats in fluid X in Fig. P2.126
such that 75 percent of its volume is submerged in fluid X.
Estimate the vacuum pressure of the air in the tank.
Air pressure?
Air = 0 kPa gage
Wood
40 cm
Fluid X
70 cm
Specific gravity
=S
M?
G
B
L
P2.128 P2.129 The iceberg idealization in Prob. P2.128 may become
­unstable if its sides melt and its height exceeds its width.
In Fig. P2.128 suppose that the height is L and the depth
into the paper is L, but the width in the plane of the paper
is H < L. Assuming S = 0.88 for the iceberg, find the
ratio H/L for which it becomes neutrally stable (about to
overturn).
P2.130 Consider a wooden cylinder (SG = 0.6) 1 m in diameter
and 0.8 m long. Would this cylinder be stable if placed to
float with its axis vertical in oil (SG = 0.8)?
P2.131 A barge is 15 ft wide and 40 ft long and floats with a
draft of 4 ft. It is piled so high with gravel that its center
of gravity is 3 ft above the waterline. Is it stable?
P2.132 A solid right circular cone has SG = 0.99 and floats
­vertically as in Fig. P2.132. Is this a stable position for
the cone?
P2.126 Water :
SG = 1.0
Stability of floating bodies
*P2.127 Consider a cylinder of specific gravity S < 1 floating
vertically in water (S = 1), as in Fig. P2.127. Derive a
formula for the stable values of D/L as a function of S
and apply it to the case D/L = 1.2.
D
L
Water
S = 1.0
h
P2.127
P2.128 An iceberg can be idealized as a cube of side length L, as
in Fig. P2.128. If seawater is denoted by S = 1.0, then
glacier ice (which forms icebergs) has S = 0.88. Determine if this “­cubic” iceberg is stable for the position
shown in Fig. P2.128.
SG = 0.99
P2.132 P2.133 Consider a uniform right circular cone of specific gravity S < 1, floating with its vertex down in water (S = 1).
The base radius is R and the cone height is H. Calculate
and plot the stability MG of this cone, in dimensionless
form, versus H/R for a range of S < 1.
P2.134 When floating in water (SG = 1.0), an equilateral triangular body (SG = 0.9) might take one of the two positions shown in Fig. P2.134. Which is the more stable
position? Assume large width into the paper.
(a)
P2.134 (b)
Problems 125
P2.135 Consider a homogeneous right circular cylinder of
length L, radius R, and specific gravity SG, floating in
water (SG = 1). Show that the body will be stable with
its axis vertical if
R
> [2SG(1 − SG) ] 1/2
L
P2.141 The same tank from Prob. P2.139 is now moving with
constant acceleration up a 30˚ inclined plane, as in Fig.
P2.141. Assuming rigid-body motion, compute (a) the
value of the acceleration a, (b) whether the acceleration
is up or down, and (c) the gage pressure at point A if the
fluid is mercury at 20˚C.
V
P2.136 Consider a homogeneous right circular cylinder of
length L, radius R, and specific gravity SG = 0.5, floating in water (SG = 1). Show that the body will be stable
with its axis horizontal if L/R > 2.0.
a?
Uniform acceleration
P2.137 A tank of water 4 m deep receives a constant upward
­acceleration az. Determine (a) the gage pressure at the
tank bottom if az = 5 m2/s and (b) the value of az that
causes the gage pressure at the tank bottom to be 1 atm.
P2.138 A 12-fl-oz glass, of 3-in diameter, partly full of water, is
attached to the edge of an 8-ft-diameter merry-go-round,
which is rotated at 12 r/min. How full can the glass be
­before water spills? Hint: Assume that the glass is much
smaller than the radius of the merry-go-round.
P2.139 The tank of liquid in Fig. P2.139 accelerates to the right
with the fluid in rigid-body motion. (a) Compute ax in
m/s2. (b) Why doesn’t the solution to part (a) depend on
the density of the fluid? (c) Determine the gage pressure
at point A if the fluid is glycerin at 20˚C.
A
100 cm
15 cm
P2.139 P2.140 The U-tube in Fig. P2.140 is moving to the right with
variable velocity. The water level in the left tube is 6 cm,
and the level in the right tube is 16 cm. Determine the
acceleration and its direction.
100 cm
28 cm
z
A
30°
x
P2.141 P2.142 The tank of water in Fig. P2.142 is 12 cm wide into the
paper. If the tank is accelerated to the right in rigid-body
motion at 6.0 m/s2, compute (a) the water depth on side
AB and (b) the water-pressure force on panel AB. Assume no spilling.
B
9 cm
Water at 20°C
A
ax
28 cm
15 cm
24 cm
P2.142
P2.143 The tank of water in Fig. P2.143 is full and open to the
atmosphere at point A. For what acceleration ax in ft/s2
will the pressure at point B be (a) atmospheric and (b)
zero absolute?
A
pa = 15 lbf/in2 abs
ax
2 ft
Water
B
1 ft
P2.140
20 cm
P2.143 1 ft
2 ft
126
Chapter 2 Pressure Distribution in a Fluid
P2.144 Consider a hollow cube of side length 22 cm, filled completely with water at 20˚C. The top surface of the cube is
horizontal. One top corner, point A, is open through a
small hole to a pressure of 1 atm. Diagonally opposite to
point A is top corner B. Determine and discuss the various rigid-body accelerations for which the water at point
B begins to cavitate, for (a) horizontal motion and (b)
vertical motion.
P2.145 A fish tank 14 in deep by 16 by 27 in is to be carried in a
car that may experience accelerations as high as 6 m/s2.
What is the maximum water depth that will avoid spilling in rigid-body motion? What is the proper alignment
of the tank with respect to the car motion?
P2.146 The tank in Fig. P2.146 is filled with water and has a vent
hole at point A. The tank is 1 m wide into the paper. Inside
the tank, a 10-cm balloon, filled with helium at 130 kPa,
is tethered centrally by a string. If the tank accelerates to
the right at 5 m/s2 in rigid-body motion, at what angle will
the balloon lean? Will it lean to the right or to the left?
60 cm
P2.148 A child is holding a string onto which is attached a helium-filled balloon. (a) The child is standing still and
suddenly accelerates forward. In a frame of reference
moving with the child, which way will the balloon tilt,
forward or backward? Explain. (b) The child is now sitting in a car that is stopped at a red light. The heliumfilled balloon is not in contact with any part of the car
(seats, ceiling, etc.) but is held in place by the string,
which is in turn held by the child. All the windows in the
car are closed. When the traffic light turns green, the car
accelerates forward. In a frame of reference moving with
the car and child, which way will the balloon tilt, forward
or backward? Explain. (c) ­Purchase or borrow a heliumfilled balloon. Conduct a scientific ­experiment to see if
your predictions in parts (a) and (b) above are correct. If
not, explain.
P2.149 The 6-ft-radius waterwheel in Fig. P2.149 is being used
to lift water with its 1-ft-diameter half-cylinder blades. If
the wheel rotates at 10 r/min and rigid-body motion is
­assumed, what is the water surface angle θ at position A?
A
1 atm
Water at 20°C
D = 10 cm
10 r/min
He
40 cm
20 cm
6 ft
String
1 ft
P2.149
P2.146
P2.147 The tank of water in Fig. P2.147 accelerates uniformly
by freely rolling down a 30˚ incline. If the wheels are
frictionless, what is the angle θ? Can you explain this
interesting result?
θ
A
P2.150 A cheap accelerometer, probably worth the price, can be
made from a U-tube as in Fig. P2.150. If L = 18 cm and
D = 5 mm, what will h be if ax = 6 m/s2? Can the scale
markings on the tube be linear multiples of ax?
D
θ
h
Rest level
ax
1
2
P2.150 30°
P2.147 1
2
L
L
L
P2.151 The U-tube in Fig. P2.151 is open at A and closed at D.
If accelerated to the right at uniform ax, what acceleration
Problems 127
will cause the pressure at point C to be atmospheric? The
fluid is water (SG = 1.0).
A
D
1 ft
1 ft
B
C
pressure is 2116 lbf/ft2 absolute, at what rotation rate will
the fluid within the tube begin to vaporize? At what point
will this occur?
P2.157 The 45˚ V-tube in Fig. P2.157 contains water and is open
at A and closed at C. What uniform rotation rate in r/min
about axis AB will cause the pressure to be equal at
points B and C? For this condition, at what point in leg
BC will the pressure be a minimum?
A
C
1 ft
P2.151
Rigid-body rotation
P2.152 A 16-cm-diameter open cylinder 27 cm high is full of
30 cm
­water. Compute the rigid-body rotation rate about its
­central axis, in r/min, (a) for which one-third of the water
will spill out and (b) for which the bottom will be barely
45°
exposed.
P2.153 A tall cylindrical container, 14 in in diameter, is used to
make a mold for forming 14-in salad bowls. The bowls
B
P2.157
are to be 8 in deep. The cylinder is half-filled with molten plastic, µ = 1.6 kg/(m-s), rotated steadily about the *P2.158 It is desired to make a 3-m-diameter parabolic telescope
central axis, then cooled while rotating. What is the apmirror by rotating molten glass in rigid-body motion unpropriate rotation rate, in r/min?
til the desired shape is achieved and then cooling the
P2.154 A very tall 10-cm-diameter vase contains 1178 cm3 of waglass to a solid. The focus of the mirror is to be 4 m from
ter. When spun steadily to achieve rigid-body rotation, a
the mirror, measured along the centerline. What is the
4-cm-diameter dry spot appears at the bottom of the vase.
proper mirror rotation rate, in r/min, for this task?
What is the rotation rate, in r/min, for this condition?
P2.159 The three-legged manometer in Fig. P2.159 is filled with
P2.155 For what uniform rotation rate in r/min about axis C will
water to a depth of 20 cm. All tubes are long and have
the U-tube in Fig. P2.155 take the configuration shown?
equal small diameters. If the system spins at angular veThe fluid is mercury at 20˚C.
locity Ω about the central tube, (a) derive a formula to
find the change of height in the tubes; (b) find the height
A
in cm in each tube if Ω = 120 r/min. Hint: The central
C
tube must supply water to both the outer legs.
B
10 cm
Ω
20 cm
10 cm
20 cm
12 cm
P2.159 Pressure measurements
P2.155
10 cm
5 cm
P2.156 Suppose that the U-tube of Fig. P2.151 is rotated about
axis DC. If the fluid is water at 122˚F and atmospheric
P2.160 Figure P2.160 shows a gage for very low pressures, invented in 1874 by Herbert McLeod. (a) Can you deduce,
from the figure, how it works? (b) If not, read about it
and explain it to the class.
128
Chapter 2 Pressure Distribution in a Fluid
P2.161 Figure P2.161 shows a sketch of a commercial pressure
gage. (a) Can you deduce, from the figure, how it works?
P2.161
P2.160
P2
P1
Word Problems
W2.1 Consider a hollow cone with a vent hole in the vertex at
the top, along with a hollow cylinder, open at the top,
with the same base area as the cone. Fill both with water
to the top. The hydrostatic paradox is that both containers have the same force on the bottom due to the water
pressure, although the cone contains 67 percent less water. Can you explain the paradox?
W2.2 Can the temperature ever rise with altitude in the real
atmosphere? Wouldn’t this cause the air pressure to increase upward? Explain the physics of this situation.
W2.3 Consider a submerged curved surface that consists of a
two-dimensional circular arc of arbitrary angle, arbitrary
depth, and arbitrary orientation. Show that the resultant hydrostatic pressure force on this surface must pass through
the center of curvature of the arc.
W2.4 Fill a glass approximately 80 percent with water, and add
a large ice cube. Mark the water level. The ice cube, having SG ≈ 0.9, sticks up out of the water. Let the ice cube
melt with negligible evaporation from the water surface.
Will the water level be higher than, lower than, or the
same as before?
W2.5 A ship, carrying a load of steel, is trapped while floating
in a small closed lock. Members of the crew want to get
W2.6
W2.7
W2.8
W2.9
out, but they can’t quite reach the top wall of the lock. A
crew member suggests throwing the steel overboard in
the lock, claiming the ship will then rise and they can
climb out. Will this plan work?
Consider a balloon of mass m floating neutrally in the
­atmosphere, carrying a person/basket of mass M > m.
­Discuss the stability of this system to disturbances.
Consider a helium balloon on a string tied to the seat
of your stationary car. The windows are closed, so
there is no air motion within the car. The car begins
to accelerate forward. Which way will the balloon
lean, forward or backward? Hint: The acceleration
sets up a horizontal pressure gradient in the air within
the car.
Repeat your analysis of Prob. W2.7 to let the car move at
constant velocity and go around a curve. Will the balloon
lean in, toward the center of curvature, or out?
The deep submersible vehicle ALVIN weighs approximately 36,000 lbf in air. It carries 800 lbm of steel
weights on the sides. After a deep mission and return,
two 400-lbm piles of steel are left on the ocean floor.
Can you explain, in terms relevant to this chapter, how
these steel weights are used?
Fundamentals of Engineering Exam Problems
FE2.1A gage attached to a pressurized nitrogen tank reads a gage
pressure of 28 in of mercury. If atmospheric pressure is
14.4 psia, what is the absolute pressure in the tank?
(a) 95 kPa, (b) 99 kPa, (c) 101 kPa, (d) 194 kPa,
(e) 203 kPa
FE2.2On a sea-level standard day, a pressure gage, moored
below the surface of the ocean (SG = 1.025), reads
an absolute pressure of 1.4 MPa. How deep is the
­instrument?
(a) 4 m, (b) 129 m, (c) 133 m, (d) 140 m, (e) 2080 m
Comprehensive Problems 129
FE2.3In Fig. FE2.3, if the oil in region B has SG = 0.8 and the
absolute pressure at point A is 1 atm, what is the absolute
pressure at point B?
(a) 5.6 kPa, (b) 10.9 kPa, (c) 107 kPa, (d) 112 kPa,
(e) 157 kPa
Oil
A
Water
SG = 1
5 cm
B
3 cm
Mercury
SG = 13.56
8 cm
4 cm
FE2.3
FE2.4In Fig. FE2.3, if the oil in region B has SG = 0.8 and the
absolute pressure at point B is 14 psia, what is the absolute pressure at point A?
(a) 11 kPa, (b) 41 kPa, (c) 86 kPa, (d) 91 kPa, (e) 101 kPa
FE2.5A tank of water (SG = 1.0) has a gate in its vertical wall 5 m
high and 3 m wide. The top edge of the gate is 2 m below the
surface. What is the hydrostatic force on the gate?
(a) 147 kN, (b) 367 kN, (c) 490 kN, (d) 661 kN,
(e) 1028 kN
FE2.6In Prob. FE2.5, how far below the surface is the center of
pressure of the hydrostatic force?
(a) 4.50 m, (b) 5.46 m, (c) 6.35 m, (d) 5.33 m,
(e) 4.96 m
FE2.7A solid 1-m-diameter sphere floats at the interface between
water (SG = 1.0) and mercury (SG = 13.56) such that 40
percent is in the water. What is the specific gravity of the
sphere?
(a) 6.02, (b) 7.28, (c) 7.78, (d) 8.54, (e) 12.56
FE2.8A 5-m-diameter balloon contains helium at 125 kPa
­absolute and 15˚C, moored in sea-level standard air. If
the gas constant of helium is 2077 m2/(s2 · K) and balloon ­material weight is neglected, what is the net lifting
force of the balloon?
(a) 67 N, (b) 134 N, (c) 522 N, (d) 653 N, (e) 787 N
FE2.9A square wooden (SG = 0.6) rod, 5 cm by 5 cm by
10 m long, floats vertically in water at 20˚C when 6
kg of steel (SG = 7.84) are attached to one end. How
high above the water surface does the wooden end of
the rod protrude?
(a) 0.6 m, (b) 1.6 m, (c) 1.9 m, (d) 2.4 m, (e) 4.0 m
FE2.10A floating body will be stable when its
(a) center of gravity is above its center of buoyancy,
(b) center of buoyancy is below the waterline, (c) center
of buoyancy is above its metacenter, (d) metacenter is
above its center of buoyancy, (e) metacenter is above its
center of gravity.
Comprehensive Problems
C2.1
Some manometers are constructed as in Fig. C2.1, where
one side is a large reservoir (diameter D) and the other
side is a small tube of diameter d, open to the atmosphere. In such a case, the height of manometer liquid on
the reservoir side does not change appreciably. This has
the advantage that only one height needs to be measured
rather than two. The manometer liquid has density ρm
while the air has density ρa. Ignore the effects of surface
tension. When there is no pressure difference across the
manometer, the elevations on both sides are the same, as
indicated by the dashed line. Height h is measured from
the zero pressure level as shown. (a) When a high pressure is applied to the left side, the manometer liquid in
the large reservoir goes down, while that in the tube at
the right goes up to conserve mass. Write an exact expression for p1gage, taking into account the movement of
the surface of the reservoir. Your equation should give
p1gage as a function of h, ρm, and the physical parameters
in the problem, h, d, D, and gravity constant g. (b) Write
an approximate expression for p1gage, neglecting the
change in elevation of the surface of the reservoir l­iquid.
(c) ­Suppose h = 0.26 m in a certain application. If pa =
101,000 Pa and the manometer liquid has a density of
820 kg/m3, estimate the ratio D/d required to keep the
error of the approximation of part (b) within 1 percent of
the exact measurement of part (a). Repeat for an error
within 0.1 percent.
To pressure measurement location
pa
D
p1
ρm
C2.1
ρa (air)
h
Zero pressure level
d
130
Chapter 2 Pressure Distribution in a Fluid
C2.2
A prankster has added oil, of specific gravity SG0, to the
left leg of the manometer in Fig. C2.2. Nevertheless, the
U-tube is still useful as a pressure-measuring device. It
is attached to a pressurized tank as shown in the figure.
(a) Find an expression for h as a function of H and other
parameters in the problem. (b) Find the special case of
your result in (a) when ptank = pa. (c) Suppose H = 5.0
cm, pa is 101.2 kPa, ptank is 1.82 kPa higher than pa, and
SG0 = 0.85. Calculate h in cm, ignoring surface tension
effects and ­neglecting air density effects.
C2.5
pa
Pressurized air tank,
with pressure = ptank
Oil
H
C2.6
h
dp ≈
Water
C2.2
C2.3
C2.4
it spills during the big drop, in which the roller coaster
achieves 0.55-g acceleration at a 45˚ angle below the horizontal. Make the calculation for him, neglecting sloshing
and ­assuming that the glass is vertical at all times.
Dry adiabatic lapse rate (DALR) is defined as the negative value of atmospheric temperature gradient, dT/dz,
when temperature and pressure vary in an isentropic
fashion. ­Assuming air is an ideal gas, DALR = –dT/dz
when T = T0(p/p0)a, where exponent a = (k – 1)/k, k = cp/
cv is the ratio of specific heats, and T0 and p0 are the
temperature and pressure at sea level, respectively. (a)
Assuming that hydrostatic conditions exist in the atmosphere, show that the dry adiabatic lapse rate is constant
and is given by DALR = g(k – 1)/(kR), where R is the
ideal gas constant for air. (b) Calculate the numerical
value of DALR for air in units of ˚C/km.
In “soft” liquids (low bulk modulus β), it may be necessary to account for liquid compressibility in hydrostatic
calculations. An approximate density relation would be
Professor F. Dynamics, riding the merry-go-round with
his son, has brought along his U-tube manometer. (You
never know when a manometer might come in handy.)
As shown in Fig. C2.3, the merry-go-round spins at constant angular velocity and the manometer legs are 7 cm
apart. The ­manometer center is 5.8 m from the axis of
rotation. ­Determine the height difference h in two ways:
(a) approximately, by assuming rigid-body translation
with a equal to the a­ verage manometer acceleration; and
(b) exactly, using rigid-body rotation theory. How good is
the approximation?
A student sneaks a glass of cola onto a roller coaster ride.
The glass is cylindrical, twice as tall as it is wide, and
filled to the brim. He wants to know what percent of the
cola he should drink before the ride begins, so that none of
C2.7
β
dρ = a2dρ
ρ
or p ≈ p0 + a2 (ρ − ρ0 )
where a is the speed of sound and (p0, ρ0) are the conditions at the liquid surface z = 0. Use this approximation to
show that the density
variation with depth in a soft liquid
2
is ρ = ρ0 e−gz/a where g is the acceleration of gravity and
z is positive upward. Then consider a vertical wall of
width b, extending from the surface (z = 0) down to depth
z = –h. Find an analytic expression for the hydrostatic
force F on this wall, and compare it with the incompressible result F = ρ0gh2b/2. Would the center of pressure be
below the incompressible position z = –2h/3?
Venice, Italy, is slowly sinking, so now, especially in
winter, plazas and walkways are flooded during storms.
The proposed solution is the floating levee of Fig. C2.7.
When filled with air, it rises to block off the sea. The
levee is 30 m high, 5 m wide, and 20 m deep. Assume a
uniform density of 300 kg/m3 when floating. For the 1-m
sea–­
lagoon difference shown, estimate the angle at
which the levee floats.
7.00 cm
Ω = 6.00 rpm
Water
h
R = 5.80 m (to center of manometer)
C2.3
Center of
rotation
Design Projects 131
C2.8
Storm levee filled
with air to float
Adriatic Sea—25 m
deep in a storm
Venice lagoon—24 m deep
Altitude z, km
0
2
5
8
Pressure p, kPa
100
78
53
34
Estimate the best-fit value of B for this data. Explain any
difficulties. [Hint: EES is recommended.]
C2.9
The ALVIN submersible vehicle has a passenger compartment which is a titanium sphere of inside diameter
78.08 in and thickness 1.93 in. If the vehicle is submerged
to a depth of 3850 m in the ocean, estimate (a) the water
pressure outside the sphere, (b) the maximum elastic
stress in the sphere, in lbf/in2, and (c) the factor of safety
of the titanium alloy (6% aluminum, 4% vanadium).
D2.2
A laboratory apparatus used in some universities is
shown in Fig. D2.2. The purpose is to measure the hydrostatic force on the flat face of the circular-arc block
and compare it with the theoretical value for given
depth h. The counterweight is arranged so that the pivot
arm is horizontal when the block is not submerged,
whence the weight W can be correlated with the hydrostatic force when the submerged arm is again brought
to horizontal. First show that the a­ pparatus concept is
valid in principle; then derive a formula for W as a
function of h in terms of the system parameters. Finally, suggest some appropriate values of Y, L, and so
on for a suitable apparatus and plot theoretical W versus
h for these values.
Hinge
Filled with water—no storm
In the U.S. Standard Atmosphere, the lapse rate B may
vary from day to day. It is not a fundamental quantity
like, say, Planck’s constant. Suppose that, on a certain
day in Rhode Island, with To = 288 K, the following
pressures are measured by weather balloons:
C2.7
Design Projects
D2.1
It is desired to have a bottom-moored, floating system that
creates a nonlinear force in the mooring line as the water
level rises. The design force F need only be accurate in the
range of seawater depths h between 6 and 8 m, as shown in
the accompanying table. Design a buoyant system that will
provide this force distribution. The system should be practical (of inexpensive materials and simple construction).
h, m
F, N
h, m
F, N
6.00
6.25
6.50
6.75
7.00
400
437
471
502
530
7.25
7.50
7.75
8.00
554
573
589
600
L
W
Counterweight
Pivot
Pivot arm
R
Side view
of block face
Fluid: ρ
h
Y
Circular arc block
D2.2
b
132
D2.3
Chapter 2 Pressure Distribution in a Fluid
The Leary Engineering Company (see Popular Science,
November 2000, p. 14) has proposed a ship hull with
hinges that allow it to open into a flatter shape when
entering shallow water. A simplified version is shown in
Fig. D2.3. In deep water, the hull cross section would be
triangular, with large draft. In shallow water, the hinges
would open to an angle as high as θ = 45˚. The dashed
Draft
D2.3
45°
line indicates that the bow and stern would be closed.
Make a parametric study of this configuration for various θ, assuming a reasonable weight and center of gravity location. Show how the draft, the metacentric height,
and the ship’s stability vary as the hinges are opened.
Comment on the effectiveness of this concept.
θ
45°
Deep water
Hinge
θ
Shallow water
References
1.
2.
3.
4.
5.
6.
7.
8.
9.
U.S. Standard Atmosphere, 1976, Government Printing
­Office, Washington, DC, 1976.
J. A. Knauss, Introduction to Physical Oceanography, 2d
ed., Waveland Press, Long Grove, IL, 2005.
E. C. Tupper, Introduction to Naval Architecture, 4th ed.,
Elsevier, New York, 2004.
E. N. Gilbert, “How Things Float,” Am. Math. Monthly, vol.
98, no. 3, 1991, pp. 201–216.
R. I. Fletcher, “The Apparent Field of Gravity in a Rotating
Fluid System,” Am. J. Phys., vol. 40, July 1972, pp. 959–965.
J. P. Holman, Experimental Methods for Engineers, 8th ed.,
McGraw-Hill, New York, 2011.
R. C. Baker, Flow Measurement Handbook, Cambridge
­University Press, New York, 2005.
T. G. Beckwith, R. G. Marangoni, and J. H. Lienhard V,
­Mechanical Measurements, 6th ed., Prentice-Hall, Upper
Saddle River, NJ, 2006.
J. W. Dally, W. F. Riley, and K. G. McConnell, Instrumentation for Engineering Measurements, 2d ed., Wiley,
New York, 1993.
10.
11.
12.
13.
14.
15.
R. J. Figliola and D. E. Beasley, Theory and Design for
­Mechanical Measurements, 4th ed., Wiley, New York,
2005.
R. W. Miller, Flow Measurement Engineering Handbook,
3d ed., McGraw-Hill, New York, 1996.
L. D. Clayton, E. P. EerNisse, R. W. Ward, and R. B.
­Wiggins, “Miniature Crystalline Quartz Electromechanical
­Structures,” Sensors and Actuators, vol. 20, Nov. 15, 1989,
pp. 171–177.
A. Kitai (ed.), Luminescent Materials and Applications,
John Wiley, New York, 2008.
B. G. Liptak (ed.), Instrument Engineer’s Handbook: Process Measurement and Analysis, 4th ed., vol. 1, CRC Press,
Boca Raton, FL, 2003.
A. von Beckerath, WIKA Handbook—Pressure and Temperature Measurement, WIKA Instrument Corp., Lawrenceville, GA, 2008.
On July 16, 1969, a massive Saturn V rocket lifted Apollo 11 from the NASA Kennedy
Space Center, carrying astronauts Neil Armstrong, Michael Collins, and Edwin Aldrin to
the first landing on the moon, four days later. The photo is filled with fluid momentum.
In this chapter we learn how to analyze both the thrust of the rocket and the force of the
exit jet on the solid surface. [NASA]
134
Chapter 3
Integral Relations
for a Control Volume
Motivation. In analyzing fluid motion, we might work with a finite region, make
a balance of flow in versus flow out, and determine gross flow effects such as
the force or torque on a body or the total energy exchange. This is referred to as
the integral approach and is the subject of this chapter.
We first develop the concept of the control volume, in nearly the same manner
as one does in a thermodynamics course, and we find the rate of change of an
arbitrary gross fluid property, a result called the Reynolds transport theorem. We
then apply this theorem, in sequence, to mass, linear momentum, angular momentum, and energy, thus deriving the basic equations of fluid mechanics in integral
form for a control volume. There are many applications, of course. The chapter
includes a special case of frictionless, s­ haft-work-free momentum and energy: the
Bernoulli equation. The Bernoulli equation is a wonderful, historic relation, but
it is extremely restrictive and should always be viewed with skepticism and care
in applying it to a real (viscous) fluid motion.
3.1 Basic Physical Laws of Fluid Mechanics
It is time now to really get serious about flow problems. The fluid statics applications of Chap. 2 were more like fun than work, at least in this writer’s opinion.
Statics problems basically require only the density of the fluid and knowledge of
the position of the free surface, but most flow problems require the analysis of
an arbitrary state of variable fluid motion defined by the geometry, the boundary
conditions, and the laws of mechanics. This chapter and the next two outline the
three basic approaches to the analysis of arbitrary flow problems:
1. Integral, or large-scale, analysis in a finite control volume (Chap. 3).
2. Differential, or small-scale, analysis in an infinitesimal control volume
(Chap. 4).
3. Experimental, or dimensional, analysis (Chap. 5).
135
136
Chapter 3 Integral Relations for a Control Volume
The three approaches are roughly equal in importance. Integral analysis, the present topic, is accurate for any flow distribution but is often based on average or
­“one-dimensional” property values at the boundaries. It always gives useful
­“engineering” estimates. In principle, the differential approach of Chap. 4 can be
applied to any problem. Only a few problems, such as straight pipe flow, yield
to exact analytical solutions. But the differential equations can be modeled numerically, and the flourishing field of computational fluid dynamics (CFD)[8] can
now be used to give good estimates for almost any geometry. Finally, the dimensional analysis of Chap. 5 applies to any problem, whether analytical, numerical,
or experimental. It is particularly useful to reduce the cost of experimentation.
Differential analysis of hydrodynamics began with Euler and d’Alembert in the
late eighteenth century. Lord Rayleigh and E. Buckingham pioneered dimensional
analysis at the end of the nineteenth century. The control volume was described
in words, on an ad hoc one-case basis, by Daniel Bernoulli in 1753. Ludwig
Prandtl, the celebrated founder of modern fluid mechanics, developed the control
volume as a systematic tool in the early 1900s. One of the writers’ teachers at
M.I.T. introduced control volume analysis into American textbooks, for thermodynamics by Keenan in 1941 [10], and for fluids by Hunsaker and Rightmire in
1947 [11]. For a complete history of the control volume, see ­Vincenti [9].
Basic Laws of Mechanics for a System
All the laws of mechanics are written for a system, which is defined as an arbitrary
quantity of mass of fixed identity. The laws of mechanics then state what happens
when there is an interaction between the system and its surroundings.
First, the system is a fixed quantity of mass, denoted by m. Thus the mass of
the system is conserved and does not change.1 This is a law of mechanics and
has a very simple mathematical form, called conservation of mass:
msyst = const
dm
= 0
dt
or
(3.1)
This is so obvious in solid mechanics problems that we often forget about it. In
fluid mechanics, we must pay a lot of attention to mass conservation.
Second, if the surroundings exert a net force F on the system, Newton’s second
law states that the mass in the system will begin to accelerate:2
F = ma = m
dV
d
= (mV)
dt
dt
(3.2)
In Eq. (2.8) we saw this relation applied to a differential element of viscous
incompressible fluid. In fluid mechanics Newton’s second law is called the linear
momentum relation. Note that it is a vector law that implies the three scalar equations Fx = max, Fy = may, and Fz = maz.
1
We are neglecting nuclear reactions, where mass can be changed to energy.
We are neglecting relativistic effects, where Newton’s law must be modified.
2
3.1 Basic Physical Laws of Fluid Mechanics 137
Third, if the surroundings exert a net moment M about the center of mass of
the system, there will be a rotation effect
dH
(3.3)
dt
where H = Σ(r × V)δm is the angular momentum of the system about its center
of mass. Here we call Eq. (3.3) the angular momentum relation. Note that it is
also a vector equation implying three scalar equations such as Mx = dHx/dt.
For an arbitrary mass and arbitrary moment, H is quite complicated and contains
nine terms (see, for example, Ref. 1). In elementary dynamics we commonly treat
only a rigid body rotating about a fixed x axis, for which Eq. (3.3) reduces to
M=
d
(ωx )
(3.4)
dt
where ωx is the angular velocity of the body and Ix is its mass moment of inertia
about the x axis. Unfortunately, fluid systems are not rigid and rarely reduce to
such a simple relation, as we shall see in Sec. 3.6.
Fourth, if heat δQ is added to the system or work δW is done by the system,
the system energy dE must change according to the energy relation, or first law
of thermodynamics:
Mx = Ix
δQ − δW = dE


dE
Q−W=
(3.5)
dt
Like mass conservation, Eq. (3.1), this is a scalar relation having only a single
component.
Finally, the second law of thermodynamics relates entropy change dS to heat
added dQ and absolute temperature T at the system boundary:
or
δQ
(3.6)
T
This is valid for a system and can be written in control volume form, but there
are almost no practical applications in fluid mechanics except to analyze flow
loss details (see Sec. 9.5).
All these laws involve thermodynamic properties, and thus we must supplement them with state relations p = p(ρ, T) and e = e(ρ, T) for the particular fluid
being studied, as in Sec. 1.8. Although thermodynamics is not the main topic of
this book, it is very important to the general study of fluid mechanics. Thermodynamics is crucial to compressible flow, Chap. 9. The student should review the
first law and the state relations, as discussed in Refs. 6 and 7.
The purpose of this chapter is to obtain mathematical expressions for these
laws valid for a control volume suitable for arbitrary regions in a flow:
dS ≥
1.
2.
3.
4.
Conservation of mass (Sec. 3.3).
The linear momentum relation (Sec. 3.4).
The angular momentum relation (Sec. 3.6).
The energy equation (Sec. 3.7).
138
Chapter 3 Integral Relations for a Control Volume
Wherever necessary to complete the analysis we also introduce a state relation
such as the perfect-gas law.
Equations (3.1) to (3.6) apply to either fluid or solid systems. They are ideal
for solid mechanics, where we follow the same system forever because it represents the product we are designing and building. For example, we follow a beam
as it deflects under load. We follow a piston as it oscillates. We follow a rocket
system all the way to Mars.
But fluid systems do not demand this concentrated attention. It is rare that we
wish to follow the ultimate path of a specific particle of fluid. Instead we wish to
know what is the influence of a fluid flow on the product that we are interested in.
For the three examples just cited, we wish to know the wind loads on the beam,
the fluid pressures on the piston, and the drag and lift loads on the rocket. This
requires that the basic laws be rewritten to apply to a specific region in the
neighborhood of our product. In other words, where the fluid particles in the wind
go after they leave the beam is of little interest to a beam designer. The user’s
point of view underlies the need for the control volume analysis of this chapter.
In analyzing a control volume, we convert the system laws to apply to a specific region, which the system may occupy for only an instant. The system passes
on, and other systems come along, but no matter. The basic laws are reformulated
to apply to this local region called a control volume. All we need to know is the
flow field in this region, and often simple assumptions will be accurate enough
(such as uniform inlet and/or outlet flows). The flow conditions away from the
control volume are then irrelevant. The technique for making such localized analyses is the subject of this chapter.
Volume and Mass Rate of Flow
All the analyses in this chapter involve evaluation of the volume flow Q or mass

flow m passing through a surface (imaginary) defined in the flow.
Suppose that the surface S in Fig. 3.1a is a sort of (imaginary) wire mesh
through which the fluid passes without resistance. How much volume of fluid
passes through S in unit time? If, typically, V varies with position, we must
integrate over the elemental surface dA in Fig. 3.1a. Also, typically V may pass
through dA at an angle θ off the normal. Let n be defined as the unit vector
normal to dA. Then the amount of fluid swept through dA in time dt is the volume
of the slanted parallelepiped in Fig. 3.1b:
d 𝒱 = V dt dA cos θ = (V · n) dA dt
The integral of d𝒱/dt is the total volume rate of flow Q through the surface S:
Q=
∫ (V · n) dA = ∫ V dA
n
s
(3.7)
s
We could replace V · n by its equivalent, Vn, the component of V normal to dA,
but the use of the dot product allows Q to have a sign to distinguish between
inflow and outflow. By convention throughout this book we consider n to be the
outward normal unit vector. Therefore V · n denotes outflow if it is positive and
3.2 The Reynolds Transport Theorem 139
Unit normal n
1
n
θ
V
θ
dA
Fig. 3.1 Volume rate of flow
through an arbitrary surface:
(a) an elemental area dA on the
surface; (b) the incremental volume
swept through dA equals V dt dA
cos θ.
V
S
dA
V dt
(a)
(b)
inflow if negative. This will be an extremely useful housekeeping device when
we are computing volume and mass flow in the basic control volume relations.

Volume flow can be multiplied by density to obtain the mass flow m. If ­density
varies over the surface, it must be part of the surface integral:

m=
∫ ρ(V · n) dA = ∫ ρV dA
n
s
s
If density and velocity are constant over the surface S, a simple expression results:

One-dimensional approximation:
m = ρQ = ρAV

Typical units for Q are m3/s and for m kg/s.
3.2 The Reynolds Transport Theorem
To convert a system analysis to a control volume analysis, we must build a bridge
that connects the mathematics between a system of individual masses and a specific region specified as the control volume. This bridge or conversion, called the
Reynolds transport theorem, can be applied to all the basic laws. Examining the
basic laws (3.1) to (3.3) and (3.5), we see that they are all concerned with the
time derivative of fluid properties m, V, H, and E. Therefore what we need is to
relate the time derivative of a system property to the rate of change of that property within a control volume.
The desired conversion formula differs slightly according to whether the
control volume is fixed, moving, or deformable. Figure 3.2 illustrates these
three cases. The fixed control volume in Fig. 3.2a encloses a stationary
region of interest to a nozzle designer. The control surface on the left slices
through the jet leaving the nozzle, encloses the surrounding atmosphere, and
slices through the flange bolts and the fluid within the nozzle. This particular
control volume exposes the stresses in the flange bolts, which contribute to
applied forces in the momentum analysis. In this sense the control volume
resembles the free-body concept, which is applied to systems in solid mechanics ­analyses.
140
Chapter 3 Integral Relations for a Control Volume
Control
surface
Control
surface
V
Fig. 3.2 Fixed, moving, and
­deformable control volumes:
(a) fixed control volume for nozzle
stress analysis; (b) control volume
moving at ship speed for drag force
analysis; (c) control volume
­deforming within cylinder for
­transient pressure variation analysis.
V
Control
surface
V
(b)
(a)
(c)
Figure 3.2b illustrates a moving control volume. Here the ship is of interest,
not the ocean, so that the control surface chases the ship at ship speed V. The
control volume is of fixed volume, but the relative motion between water and
ship must be considered. If V is constant, this relative motion is a steady flow
pattern, which simplifies the analysis.3 If V is variable, the relative motion is
unsteady, so that the computed results are time-variable and certain terms enter
the momentum analysis to reflect the noninertial (accelerating) frame of reference.
Figure 3.2c shows a deforming control volume. Varying relative motion at the
boundaries becomes a factor, and the rate of change of shape of the control volume enters the analysis. We begin by deriving the fixed control volume case, and
we consider the other cases as advanced topics. An interesting history of control
volume analysis is given by Vincenti [9].
Arbitrary Fixed Control Volume
Figure 3.3 shows a fixed control volume with an arbitrary flow pattern passing
through. There are variable slivers of inflow and outflow of fluid all about the
control surface. In general, each differential area dA of surface will have a different velocity V making a different angle θ with the local normal to dA. Some
elemental areas will have inflow volume (VA cos θ)in dt, and others will have
outflow volume (VA cos θ)out dt, as seen in Fig. 3.3. Some surfaces might correspond to streamlines (θ = 90°) or solid walls (V = 0) with neither inflow nor
outflow.
Let B be any property of the fluid (energy, momentum, enthalpy, etc.) and let
β = dB/dm be the intensive value, or the amount of B per unit mass in any small
element of the fluid. The total amount of B in the control volume (the solid curve
in Fig. 3.3) is thus
BCV =
∫
CV
β dm =
∫
CV
βρ d 𝒱
β=
dB
dm
(3.8)
3
A wind tunnel uses a fixed model to simulate flow over a body moving through a fluid. A tow
tank uses a moving model to simulate the same situation.
3.2 The Reynolds Transport Theorem 141
System at
time t + dt
Vout
System at
time t
θ
n, Unit outward
normal to dA
dA
Fixed
control
volume
CV
θ
dA
Vin
n, Unit outward
normal to d A
Fig. 3.3 An arbitrary control volume
with an arbitrary flow pattern.
Arbitrary
fixed
control
surface
CS
d𝒱out = Vout d A out cos θ out dt
= V ∙ n dA dt
d𝒱in = Vin d Ain cos θ in dt
= –V ∙ n d A dt
Examining Fig. 3.3, we see three sources of changes in B relating to the control volume:
d
dt (
A change within the control volume
Outflow of β from the control volume
Inflow of β to the control volume
∫
CS
∫
CV
∫
CS
β ρd 𝒱)
βρV cos θ dAout
(3.9)
βρV cos θ dAin
The notations CV and CS refer to the control volume and control surface,
respectively. Note, in Fig. 3.3, that the system has moved a bit. In the limit as
dt → 0, the instantaneous change of B in the system is the sum of the change
within, plus the outflow, minus the inflow:
d
d
(Bsyst ) = (
dt
dt
∫
CV
βρ d 𝒱) +
∫
CS
βρV cos θ dAout −
∫
CS
βρV cos θ dAin
(3.10)
This is the Reynolds transport theorem for an arbitrary fixed control volume. By
letting the property B be mass, momentum, angular momentum, or energy, we
can rewrite all the basic laws in control volume form. Note that all three of the
integrals are concerned with the intensive property β. Since the control volume
is fixed in space, the elemental volumes d 𝒱 do not vary with time, so that the
time derivative of the volume integral vanishes unless either β or ρ varies with
time (unsteady flow).
142
Chapter 3 Integral Relations for a Control Volume
Equation (3.10) expresses the basic formula that a system derivative equals
the rate of change of B within the control volume plus the flow of B out of
the control surface minus the flow of B into the control surface. The quantity
B (or β) may be any vector or scalar property of the fluid. Two alternate forms
are possible for the flow terms. First we may notice that V cos θ is the component of V normal to the area element of the control surface. Thus we can
write
Flow terms =
∫
CS
βρVn dAout −
∫
CS
βρVn dAin =
∫
CS

β dmout −
∫
CS

β dmin
(3.10a)

where dm = ρVn dA is the differential mass flow through the surface. Form
(3.10a) helps us visualize what is being calculated.
A second, alternative form offers elegance and compactness as advantages. If
n is defined as the outward normal unit vector everywhere on the control surface,
then V · n = Vn for outflow and V · n = −Vn for inflow. Therefore the flow
terms can be represented by a single integral involving V · n that accounts for
both positive outflow and negative inflow:
Flow terms =
∫
βρ(V · n) dA
CS
(3.11)
The compact form of the Reynolds transport theorem is thus
d
d
(Bsyst ) = (
dt
dt
∫
CV
βρ d 𝒱) +
∫
CS
βρ(V · n) dA (3.12)
This is beautiful but only occasionally useful, when the coordinate system is
ideally suited to the control volume selected. Otherwise the computations are
easier when the flow of B out is added and the flow of B in is subtracted,
according to Eqs. (3.10) or (3.11).
The time derivative term can be written in the equivalent form
d
dt (
∫
CV
βρ d𝒱) =
∫
CV
∂
(βρ) d𝒱
∂t
(3.13)
for the fixed control volume since the volume elements do not vary.
Control Volume Moving at Constant Velocity
If the control volume is moving uniformly at velocity Vs, as in Fig. 3.2b, an
observer fixed to the control volume will see a relative velocity Vr of fluid crossing the control surface, defined by
Vr = V − Vs (3.14)
where V is the fluid velocity relative to the same coordinate system in which the
control volume motion Vs is observed. Note that Eq. (3.14) is a vector subtraction.
The flow terms will be proportional to Vr, but the volume integral of Eq. (3.12)
is unchanged because the control volume moves as a fixed shape without
3.2 The Reynolds Transport Theorem 143
d­ eforming. The Reynolds transport theorem for this case of a uniformly moving
control volume is
d
d
(Bsyst ) = (
dt
dt
∫
CV
βρ d 𝒱) +
∫
βρ(Vr · n) dA
CS
(3.15)
which reduces to Eq. (3.12) if Vs ≡ 0.
Control Volume of Constant Shape but Variable Velocity4
If the control volume moves with a velocity Vs(t) that retains its shape, then the
volume elements do not change with time, but the boundary relative velocity
Vr = V(r, t) − Vs(t) becomes a somewhat more complicated function. Equation
(3.15) is unchanged in form, but the area integral may be more laborious to
evaluate.
Arbitrarily Moving and Deformable Control Volume5
The most general situation is when the control volume is both moving and
deforming arbitrarily, as illustrated in Fig. 3.4. The flow of volume across the
control surface is again proportional to the relative normal velocity component
System at
time t + dt
CV at time t + dt
System and
CV at time t
V
Vs
Vr
Vs
Vr =
V
V – Vs
n
n
d𝒱out = ( Vr ∙ n) d A dt
d𝒱in = –(Vr ∙ n) d A d t
Fig. 3.4 Relative velocity effects ­between a system and a control ­volume
when both move and ­deform. The system boundaries move at velocity V,
and the control surface moves at velocity Vs.
4
5
This section may be omitted without loss of continuity.
This section may be omitted without loss of continuity.
144
Chapter 3 Integral Relations for a Control Volume
Vr · n, as in Eq. (3.15). However, since the control surface has a deformation, its
velocity Vs = Vs(r, t), so that the relative velocity Vr = V(r, t) − Vs(r, t) is or
can be a complicated function, even though the flow integral is the same as in
Eq. (3.15). Meanwhile, the volume integral in Eq. (3.15) must allow the volume
elements to distort with time. Thus the time derivative must be applied after
integration. For the deforming control volume, then, the transport theorem takes
the form
d
d
(Bsyst ) = (
dt
dt
∫
CV
βρ d 𝒱) +
∫
CS
βρ(Vr · n) dA (3.16)
This is the most general case, which we can compare with the equivalent form
for a fixed control volume:
d
(Bsyst ) =
dt
∫
CV
∂
( βρ) d𝒱 +
∂t
∫
βρ(V · n) dA
CS
(3.17)
The moving and deforming control volume, Eq. (3.16), contains only two complications: (1) The time derivative of the first integral on the right must be taken
outside, and (2) the second integral involves the relative velocity Vr between the
fluid system and the control surface. These differences and mathematical subtleties are best shown by examples.
One-Dimensional Flux Term Approximations
In many situations, the flow crosses the boundaries of the control surface only at
simplified inlets and exits that are approximately one-dimensional; that is, flow
properties are nearly uniform over the cross section. For a fixed control volume,
the surface integral in Eq. (3.12) reduces to a sum of positive (outlet) and negative (inlet) product terms for each cross section:
d
d
(Bsyst ) = (
dt
dt
∫
CV
β dm) +
∑ βi m i ∣ out − ∑ βim i ∣ in where m i = ρiAiVi outlets
inlets
(3.18)
To the writer, this is an attractive way to set up a control volume analysis without
using the dot product notation. An example of multiple one-dimensional flows is
shown in Fig. 3.5. There are inlet flows at sections 1 and 4 and outflows at
­sections 2, 3, and 5. Equation (3.18) becomes
d
d
(Bsyst ) = (
dt
dt
∫
CV
β dm) + β2 (ρAV) 2 + β3 (ρAV) 3 + β5 (ρAV) 5
− β1 (ρAV) 1 − β4 (ρAV) 4
(3.19)
with no contribution from any other portion of the control surface because there
is no flow across the boundary.
3.2 The Reynolds Transport Theorem 145
Section 2:
uniform V2 , A2 , ρ 2 , β 2 , etc.
CS
2
3
1
All sections i:
Vi approximately
normal to area Ai
CV
4
Fig. 3.5 A control volume with
simplified one-dimensional inlets
and exits.
5
EXAMPLE 3.1
A fixed control volume has three one-dimensional boundary sections, as shown in
Fig. E3.1. The flow within the control volume is steady. The flow properties at each
section are tabulated below. Find the rate of change of energy of the system that
­occupies the control volume at this instant.
3
CV
1
E3.1
2
Section
Type
ρ, kg/m3
V, m/s
A, m2
e, J/kg
1
2
3
Inlet
Inlet
Outlet
800
800
800
5.0
8.0
17.0
2.0
3.0
2.0
300
100
150
Solution
∙ System sketch: Figure E3.1 shows two inlet flows, 1 and 2, and a single outlet flow, 3.
∙ Assumptions: Steady flow, fixed control volume, one-dimensional inlet and exit flows.
∙ Approach: Apply Eq. (3.17) with energy as the property, where B = E and
β = dE/dm = e. Use the one-dimensional flow approximation and then insert the
data from the table.
∙ Solution steps: Outlet 3 contributes a positive term, and inlets 1 and 2 are negative.
The appropriate form of Eq. (3.12) is
dE
d
( dt )syst = dt (
#
CV



e ρ d υ) + e3 m3 − e1 m1 − e2 m2
146
Chapter 3 Integral Relations for a Control Volume
Since the flow is steady, the time-derivative volume integral term is zero. Introducing
(ρAV)i as the mass flow grouping, we obtain
dE
( dt )syst = −e1ρ1A1V1 − e2ρ2A2V2 + e3ρ3A3V3
Introducing the numerical values from the table, we have
dE
3
2
( dt )syst = −(300 J/kg) (800 kg/m ) (2 m ) (5 m/s) − 100(800) (3) (8) + 150(800) (2) (17)
= (−2,400,000 − 1,920,000 + 4,080,000) J/s
= −240,000 J/s = −0.24 MJ/s
Ans.
Thus the system is losing energy at the rate of 0.24 MJ/s = 0.24 MW. Since we have
accounted for all fluid energy crossing the boundary, we conclude from the first law
that there must be heat loss through the control surface, or the system must be doing
work on the environment through some device not shown. Notice that the use of SI
units leads to a consistent result in joules per second without any conversion factors.
We promised in Chap. 1 that this would be the case.
∙ Comments: This problem involves energy, but suppose we check the balance of
mass also. Then B = mass m, and β = dm/dm = unity. Again the volume integral
vanishes for steady flow, and Eq. (3.17) reduces to
dm
( dt )syst =
∫
CS
ρ(V · n) dA = −ρ1A1V1 − ρ2A2V2 + ρ3 A3V3
= −(800 kg/m3 ) (2 m2 ) (5 m/s) − 800(3) (8) + 800(17) (2)
= (−8000 − 19,200 + 27,200) kg/s = 0 kg/s
Thus the system mass does not change, which correctly expresses the law of conservation of system mass, Eq. (3.1).
EXAMPLE 3.2
Compressed air in a rigid tank of volume 𝒱 exhausts through a small nozzle as in
Fig. E3.2. Air properties change through the nozzle, and the flow exits at ρo,Vo, Ao.
Find an expression for the rate of change of tank density.
CV
ρ, p, T
E3.2
ρ o, po, Vo, Ao
Solution
∙ System sketch: Fig. E3.2 shows one exit, no inlets. The constant exit area is Ao.
∙ Control volume: As shown, we choose a CV that encircles the entire tank and
nozzle.
∙ Assumptions: Unsteady flow (the tank mass decreases), one-dimensional exit flow.
∙ Approach: Apply Eq. (3.16) for mass, B = m and β = dm/dm = unity.
∙ Solution steps: Write out the Reynolds transport relation (3.16) for this problem:
dm
d
( dt )syst = 0 = dt (
∫
ρ d 𝒱) +
CV
∫
CS
ρ(V · n) dA = 𝒱
dρ
+ ρo Vo Ao
dt
3.3 Conservation of Mass 147
Solve for the rate of change of tank density:
dρ
ρ o Vo Ao
=−
dt
𝒱
Ans.
∙ Comments: This is a first-order ordinary differential equation for the tank density.
If we account for changes in ρo and Vo from the compressible-flow theories of
Chap. 9, we can readily solve this equation for the tank density ρ(t).
For advanced study, many more details of the analysis of deformable control
volumes can be found in Hansen [4] and Potter et al. [5].
3.3 Conservation of Mass
The Reynolds transport theorem, Eq. (3.16) or (3.17), establishes a relation
between system rates of change and control volume surface and volume integrals. But system derivatives are related to the basic laws of mechanics, Eqs.
(3.1) to (3.5). Eliminating system derivatives between the two gives the control
volume, or integral, forms of the laws of mechanics of fluids. The dummy variable B becomes, respectively, mass, linear momentum, angular momentum, and
energy.
For conservation of mass, as discussed in Examples 3.1 and 3.2, B = m and
β = dm/dm = 1. Equation (3.1) becomes
dm
d
( dt )syst = 0 = dt (
∫
CV
ρ d 𝒱) +
∫
ρ(Vr · n) dA
CS
(3.20)
This is the integral mass conservation law for a deformable control volume. For
a fixed control volume, we have
∫
CV
∂ρ
d𝒱 +
∂t
∫
CS
ρ(V · n) dA = 0
(3.21)
If the control volume has only a number of one-dimensional inlets and outlets,
we can write
∫
CV
∂ρ
d𝒱 +
∂t
∑ (ρi Ai Vi ) out − ∑ (ρi Ai Vi ) in = 0 i
i
(3.22)
Other special cases occur. Suppose that the flow within the control volume is
steady; then ∂ρ/∂t ≡ 0, and Eq. (3.21) reduces to
∫
CS
ρ(V · n) dA = 0
(3.23)
148
Chapter 3 Integral Relations for a Control Volume
This states that in steady flow the mass flows entering and leaving the control
volume must balance exactly.6 If, further, the inlets and outlets are one-dimensional,
we have for steady flow
∑ (ρi Ai Vi ) in = ∑ (ρi A i Vi ) out
i
i
(3.24)
This simple approximation is widely used in engineering analyses. For example,
referring to Fig. 3.5, we see that if the flow in that control volume is steady, the
three outlet mass flows balance the two inlets:
Outflow = Inflow
ρ2A2V2 + ρ3A3V3 + ρ5A5V5 = ρ1A1V1 + ρ4A4V4
(3.25)

The quantity ρAV is called the mass flow m passing through the one-dimensional
cross section and has consistent units of kilograms per second (or slugs per second) for SI (or BG) units. Equation (3.25) can be rewritten in the short form





m2 + m3 + m5 = m1 + m4
(3.26)
and, in general, the steady-flow–mass-conservation relation (3.23) can be written as
∑ (m i ) out = ∑ (m i ) in
i
i
(3.27)

If the inlets and outlets are not one-dimensional, one has to compute m by integration over the section

mcs =
∫
cs
ρ(V · n) dA
(3.28)
where “cs” stands for cross section. An illustration of this is given in Example
3.4.
Incompressible Flow
Still further simplification is possible if the fluid is incompressible, which we
may define as having density variations that are negligible in the mass conservation requirement.7 As we saw in Chap. 1, all liquids are nearly incompressible,
and gas flows can behave as if they were incompressible, particularly if the gas
velocity is less than about 30 percent of the speed of sound of the gas.
6
Throughout this section we are neglecting sources or sinks of mass that might be embedded in
the control volume. Equations (3.20) and (3.21) can readily be modified to add source and sink
terms, but this is rarely necessary.
7
Be warned that there is subjectivity in specifying incompressibility. Oceanographers consider a
0.1 percent density variation very significant, while aerodynamicists may neglect density variations
in highly compressible, even hypersonic, gas flows. Your task is to justify the incompressible
approximation when you make it.
3.3 Conservation of Mass 149
Again consider the fixed control volume. For nearly incompressible flow, the term
∂ρ/∂t is small, so the time-derivative volume integral in Eq. (3.21) can be neglected.
The constant density can then be removed from the surface integral for a nice
simplification:
d
dt (
∫
CV
∂ρ
dυ) +
∂t
∫
or
∫
ρ(V · n) dA = 0
CS
(V · n) dA = 0
(3.29)
CS
If the inlets and outlets are one-dimensional, we have
∑ (Vi Ai ) out = ∑ (Vi Ai ) in
i
i
or
(3.30)
∑Qout = ∑Qin
where Qi = ViAi is called the volume flow passing through the given cross
section.
Again, if consistent units are used, Q = VA will have units of cubic meters
per second (SI) or cubic feet per second (BG). If the cross section is not onedimensional, we have to integrate
QCS =
∫
CS
(V · n) dA
(3.31)
Equation (3.31) allows us to define an average velocity Vav that, when multiplied
by the section area, gives the correct volume flow:
∫
Q 1
=
(V · n) dA
A A
Vav =
(3.32)
This could be called the volume-average velocity. If the density varies across the
section, we can define an average density in the same manner:
ρav =
∫
1
ρ dA
A
(3.33)
But the mass flow would contain the product of density and velocity, and the
average product (ρV)av would in general have a different value from the product
of the averages:
(ρV) av =
∫
1
ρ(V · n) dA ≈ ρavVav
A
(3.34)
We illustrate average velocity in Example 3.4. We can often neglect the difference or, if necessary, use a correction factor between mass average and volume
average.
Chapter 3 Integral Relations for a Control Volume
150
EXAMPLE 3.3
Write the conservation-of-mass relation for steady flow through a streamtube (flow everywhere parallel to the walls) with a single one-dimensional inlet 1 and exit 2 (Fig. E3.3).
V∙n=0
V2
2
V1
1
Streamtube
control volume
Solution
For steady flow Eq. (3.24) applies with the single inlet and exit:

m = ρ1A1V1 = ρ2A2V2 = const
Thus, in a streamtube in steady flow, the mass flow is constant across every section of
the tube. If the density is constant, then
E3.3
Q = A1V1 = A2V2 = const
or
V2 =
A1
V1
A2
The volume flow is constant in the tube in steady incompressible flow, and the velocity increases as the section area decreases. This relation was derived by Leonardo da
Vinci in 1500.
EXAMPLE 3.4
For steady viscous flow through a circular tube (Fig. E3.4), the axial velocity profile
is given approximately by
r=R
r
x
u = 0 (no slip)
E3.4
u(r)
u = U0 (1 −
r m
R)
U0
so that u varies from zero at the wall (r = R), or no slip, up to a maximum u = U0 at
the centerline r = 0. For highly viscous (laminar) flow m ≈ 12 , while for less viscous
(turbulent) flow m ≈ 17 . Compute the average velocity if the density is constant.
Solution
The average velocity is defined by Eq. (3.32). Here V = iu and n = i, and thus V · n
= u. Since the flow is symmetric, the differential area can be taken as a circular strip
dA = 2 πr dr. Equation (3.32) becomes
Vav =
or
∫
1
1
u dA = 2
A
πR
Vav = U0
∫ U (1 − Rr ) 2πr dr
R
0
m
0
2
(1 + m) (2 + m)
Ans.
For the laminar flow approximation, m ≈ 12 and Vav ≈ 0.53U0. (The exact laminar
theory in Chap. 6 gives Vav = 0.50U0.) For turbulent flow, m ≈ 17 and Vav ≈ 0.82U0.
(There is no exact turbulent theory, and so we accept this a­ pproximation.) The turbulent
velocity profile is more uniform across the section, and thus the average velocity is
only slightly less than maximum.
3.3 Conservation of Mass 151
Tank area A t
EXAMPLE 3.5
The tank in Fig. E3.5 is being filled with water by two one-dimensional inlets. Air is
trapped at the top of the tank. The water height is h. (a) Find an expression for the
change in water height dh/dt. (b) Compute dh/dt if D1 = 1 in, D2 = 3 in, V1 = 3 ft/s,
V2 = 2 ft/s, and At = 2 ft2, assuming water at 20°C.
ρa
H
h
ρw
2
1
Fixed CS
Solution
Part (a)
A suggested control volume encircles the tank and cuts through the two inlets. The flow
within is unsteady, and Eq. (3.22) applies with no outlets and two inlets:
E3.5
d
dt (
∫
CV
ρ d 𝒱) − ρ1A1V1 − ρ2A2V2 = 0
(1)
Now if At is the tank cross-sectional area, the unsteady term can be evaluated as
follows:
d
dt (
∫
CV
ρ d 𝒱) =
d
d
dh
(ρw Ath) + [ρa At (H − h) ] = ρw At dt
dt
dt
(2)
The ρa term vanishes because it is the rate of change of air mass and is zero because
the air is trapped at the top. Substituting (2) into (1), we find the change of water height
dh ρ1A1V1 + ρ2A2V2
=
dt
ρ w At
Ans. (a)
For water, ρ1 = ρ2 = ρw, and this result reduces to
dh A1V1 + A2V2 Q1 + Q2
=
=
dt
At
At
(3)
Part (b)
The two inlet volume flows are
Q1 = A1V1 = 14π( 121 ft) 2 (3 ft/s) = 0.016 ft3/s
Q2 = A2V2 = 14π( 123 ft) 2 (2 ft/s) = 0.098 ft3/s
Then, from Eq. (3),
dh (0.016 + 0.098) ft3/s
=
= 0.057 ft/s
dt
2 ft2
Ans. (b)
Suggestion: Repeat this problem with the top of the tank open.
The control volume mass relations, Eq. (3.20) or (3.21), are fundamental to
all fluid flow analyses. They involve only velocity and density. Vector directions
are of no consequence except to determine the normal velocity at the surface and
hence whether the flow is in or out. Although your specific analysis may concern
152
Chapter 3 Integral Relations for a Control Volume
forces or moments or energy, you must always make sure that mass is balanced
as part of the analysis; otherwise the results will be unrealistic and probably
incorrect. We shall see in the examples that follow how mass conservation is
constantly checked in performing an analysis of other fluid properties.
3.4 The Linear Momentum Equation
In Newton’s second law, Eq. (3.2), the property being differentiated is the linear
momentum mV. Therefore our dummy variable is B = mV and β = dB/dm = V,
and application of the Reynolds transport theorem gives the linear momentum
relation for a deformable control volume:
d
(mV) syst =
dt
∑F =
d
dt (
∫
Vρ d 𝒱) +
CV
∫
Vρ(Vr · n) dA (3.35)
CS
The following points concerning this relation should be strongly emphasized:
1. The term V is the fluid velocity relative to an inertial (nonaccelerating)
coordinate system; otherwise Newton’s second law must be modified to
include noninertial relative acceleration terms (see the end of this section).
2. The term Σ F is the vector sum of all forces acting on the system material
considered as a free body; that is, it includes surface forces on all fluids
and solids cut by the control surface plus all body forces (gravity and electromagnetic) acting on the masses within the control volume.
3. The entire equation is a vector relation; both the integrals are vectors due to
the term V in the integrands. The equation thus has three components. If we
want only, say, the x component, the equation reduces to
∑ Fx =
d
dt (
∫
CV
uρ d 𝒱) +
∫
uρ(Vr · n) dA
CS
(3.36)
and, similarly, Σ Fy and Σ Fz would involve v and w, respectively. Failure to
account for the vector nature of the linear momentum relation (3.35) is probably the greatest source of student error in control volume analyses.
For a fixed control volume, the relative velocity Vr ≡ V, and Eq. (3.35)
becomes
∑F =
d
dt (
∫
Vρ d 𝒱) +
CV
∫
Vρ(V · n) dA
CS
(3.37)
Again we stress that this is a vector relation and that V must be an inertial-frame
velocity. While most of the momentum analyses in this text are concerned with
Eq. (3.37), it is useful to present several important generalities for its applications.
They are demonstrated by example problems to illustrate features of various forms
3.4 The Linear Momentum Equation 153
of Eq. (3.37) for a control volume. We describe these generalities under subheadings mostly with examples. Example 3.6 explains how to simplify pressure force
on a closed control volume by working in gage pressures. Examples 3.7 and 3.8
demonstrate how choice of the control volume can help analysis of external
anchoring forces. Example 3.9 shows how choice of the control volume can
simplify analysis of friction or drag force. Example 3.10 explains how to set up
pressure conditions for flows at a sluice gate. Example 3.11 involves a problem
with a moving object. Finally we will summarize a few tips used in the analysis
of the linear momentum equation.
One-Dimensional Momentum Flux
By analogy with the term mass flow used in Eq. (3.28), the surface integral in
Eq. (3.37) is called the momentum flow term. If we denote momentum by M, then

M CS =
∫
Vρ(V · n) dA
sec
(3.38)
Because of the dot product, the result will be negative for inlet momentum flow
and positive for outlet flow. If the cross section is one-dimensional, V and ρ are
uniform over the area and the integrated result is


M seci = Vi (ρiVni Ai ) = miVi
(3.39)

for outlet flow and −miVi for inlet flow. Thus if the control volume has only
one-dimensional inlets and outlets, Eq. (3.37) reduces to
∑F =
d
dt (
∫


Vρ d 𝒱) + ∑ (miVi ) out − ∑ (miVi ) in
CV
(3.40)
This is a commonly used approximation in engineering analyses. It is crucial to
realize that we are dealing with vector sums. Equation (3.40) states that the net
vector force on a fixed control volume equals the rate of change of vector momentum within the control volume plus the vector sum of outlet momentum flows
minus the vector sum of inlet flows.
One-dimensional flows are easier to work with than flows involving nonuniform velocity distributions. For viscous flow through a circular pipe or duct, the
axial velocity is usually nonuniform as seen in Example 3.4. But we can use the
average velocity of the pipe in the momentum flux calculation with a correction
factor to be introduced in this section.
Note that the time rate of change of the vector momentum within a nondeforming control volume is zero for steady flow. The linear momentum problems
­considered in this text are all steady flow.
Net Pressure Force on a Closed Control Surface
Generally speaking, the surface forces on a control volume are due to (1) forces
exposed by cutting through solid bodies that protrude through the surface and
(2) forces due to pressure and viscous stresses of the surrounding fluid. The
154
Chapter 3 Integral Relations for a Control Volume
pgage = p – pa
n
n
pa
pa
pa
pa
Closed
CS
Fig. 3.6 Pressure force computation
by subtracting a uniform
­distribution: (a) uniform pressure,
pa
∫
Closed
CS
pa
F = −pa n dA ≡ 0;
pgage
pgage
(b) nonuniform pressure,
∫
F = − ( p − pa )n dA.
pgage = 0
(a)
(b)
computation of pressure force is relatively simple, as shown in Fig. 3.6. Recall
from Chap. 2 that the external pressure force on a surface is normal to the surface
and inward. Since the unit vector n is defined as outward, one way to write the
pressure force is
Fpress =
∫
(3.41)
p(−n) dA
CS
Now if the pressure has a uniform value pa all around the surface, as in Fig. 3.7a,
the net pressure force is zero:
FUP =
∫ p (−n)dA = −p ∫ n dA ≡ 0
a
a
(3.42)
where the subscript UP stands for uniform pressure. This result is independent of the
shape of the surface8 as long as the surface is closed and all our control volumes are
closed. Thus a seemingly complicated pressure force problem can be simplified by
subtracting atmospheric pressure pa and working only with the pieces of gage pressure
that remain, as illustrated in Fig. 3.6b. So Eq. (3.41) is entirely equivalent to
Fpress =
∫
CS
(p − pa )(−n)dA =
∫
CS
pgage (−n)dA
This trick can mean quite a savings in computation.
EXAMPLE 3.6
A control volume of a nozzle section has surface pressures of 40 lbf/in2 absolute at
section 1 and atmospheric pressure of 15 lbf/in2 absolute at section 2 and on the external rounded part of the nozzle, as in Fig. E3.6a. Compute the net pressure force if
D1 = 3 in and D2 = 1 in.
8
Can you prove this? It is a consequence of Gauss’s theorem from vector analysis.
3.4 The Linear Momentum Equation 155
Solution
∙ System sketch: The control volume is the outside of the nozzle, plus the cut sections
(1) and (2). There would also be stresses in the cut nozzle wall at section 1, which
we are neglecting here. The pressures acting on the control volume are shown in Fig.
E3.6a. Figure E3.6b shows the pressures after 15 lbf/in2 has been subtracted from all
sides. Here we compute the net pressure force only.
Jet exit pressure is atmospheric
25 lbf/in2 gage
15 lbf/in2 abs
40 lbf/in2 abs
0 lbf/in2 gage
15 lbf/in2
abs
Flow
0 lbf/in2 gage
Flow
2
2
1
E3.6
15 lbf/in2 abs
(a)
0
lbf/in2
gage
1
(b)
∙ Assumptions: Known pressures, as shown, on all surfaces of the control volume.
∙ Approach: Since three surfaces have p = 15 lbf/in2, subtract this amount everywhere
so that these three sides reduce to zero “gage pressure” for convenience. This is
allowable because of Eq. (3.42).
∙ Solution steps: For the modified pressure distribution, Fig. E3.6b, only section 1 is
needed:
Fpress = pgage,1 (−n) 1 A1 = (25
lbf
π
−(−i) ][ (3 in) 2 ] = 177i lbf
4
in2 )[
Ans.
∙ Comments: This “uniform subtraction” artifice, which is entirely legal, has greatly
simplified the calculation of pressure force. Note: We were a bit too informal when
multiplying pressure in lbf/in2 times area in square inches. We achieved lbf correctly,
but it would be better practice to convert all data to standard BG units. Further note:
In addition to Fpress, there are other forces involved in this flow, due to tension stresses
in the cut nozzle wall and the fluid weight inside the control volume.
Pressure Condition at a Jet Exit
Figure E3.6 illustrates a pressure boundary condition commonly used for jet exit
flow problems. When a fluid flow leaves a confined internal duct and exits into
an ambient “atmosphere,” its free surface is exposed to that atmosphere. Therefore
the jet itself will essentially be at atmospheric pressure also. This condition was
used in Fig. E3.6.
Only two effects could maintain a pressure difference between the atmosphere
and a free exit jet. The first is surface tension, Eq. (1.31), which is usually negligible. The second effect is a supersonic jet, which can separate itself from an
156
Chapter 3 Integral Relations for a Control Volume
atmosphere with expansion or compression waves (Chap. 9). For the majority of
applications, therefore, we shall set the pressure in an exit jet as atmospheric.
Anchoring Force to Hold an Elbow or Vane Stationary
When a liquid flow passes through an elbow that changes the direction of flow
in a piping system, or water jetting out from a nozzle is deflected by a vane into
different direction, it is crucial to determine the anchoring forces required to hold
the elbow and vane stationary.
EXAMPLE 3.7
Water flows steadily through a section of expanding elbow shown in the diagram. At
the inlet to the elbow, the absolute pressure is 230 kPa and the cross-sectional area is
0.004 m2. The direction of flow is 45° from horizontal. At the outlet, the absolute
pressure is 200 kPa, the cross-sectional area is 0.01 m2, and the velocity is 2 m/s.
Given water density of 1000 kg/m3, determine the anchoring forces in x and y directions
required to hold the elbow in place.
Solution
Choose a fixed control volume in Fig. 3.7a that includes the elbow, cuts through the
inlet and exit of the elbow, and includes the elbow support (not shown in the figure).
The friction between the flow and the inner surface of the elbow is internally selfcanceling. If we subtract Patm from the entire control surface, we can work in gage
pressures at the inlet and exit of the elbow. If we assume uniform flow at the elbow
inlet and exit, and atmospheric pressure Patm = 101 kPa, and neglect weight of elbow
and water in elbow, then Eq. (3.40) reduces to
∑ F = m 2V2 − m 1V1(1)



Conservation of mass for the flow requires m1 = m2 = m = ρAV . Referring to the coordinates in Fig. 3.7(a), we can readily find the force components in x and y directions,
respectively.
In x direction,
Rx + P1, gage A1 cos θ − p2, gage A2 = (ρ2A2V2 )V2 − (ρ1A1V1 )V1 cos θ(2)
V∙n=0
V2
.
m = constant
P2 A2
Fig. 3.7 Net forces on an expanding
elbow: (a) geometry of the elbow
turning the steady water flow;
(b) vector diagram for the forces and
momentum.
P1 A1
θ
V1
Fixed
control
volume
1
(a)
.
m V1
2
y
x
.
ΣF = m (V2 – V1)
θ
.
m V2
(b)
3.4 The Linear Momentum Equation 157
For incompressible flow, ρ = const.
V 1 = V2
A2
0.01 m2
= 2 m∕s
= 5 m∕s
A1
0.004 m2
Rearranging Eq. (1), it becomes
Rx = −P1, gage A1 cos θ + p2, gage A2 + (ρ2A2V2 )V2 − (ρ1A1V1 )V1 cos θ(3)
For the given numerical values, we have
Rx = −(129 kPa)(0.004 m2) cos 45 + (99 kPa)(0.01 m2) + 1000
−(0.004 m2) 52
m2
cos 45] = 594.4 N
s2
2
2
2m
(0.01
m
)
2
m3 [
s2
kg
Ans.
In y direction,
Ry + P1, gage A1 sin θ = 0 − (ρ1A1V1 )V1 sin θ(4)
Rearranging Eq. (4), it becomes
Ry = −P1, gage A1 sin θ = 0 − (ρ1A1V1 )V1 sin θ(5)
For the given numerical values, we have
Ry = −(129 kPa)(0.004 m2) sin 45 −1000
kg
m3
(0.004 m2) 52
m2
sin 45 = −435.6 N Ans.
s2
Negative sign of Ry indicates that the anchoring force component is in the direction
facing downward.
Selecting an appropriate control volume that contains the elbow and water in the
elbow leads directly to the solution for the anchoring force. Use of the gage pressures
at the inlet and exit simplifies evaluation of the surface forces. The weight of the elbow
and of the water in the elbow are neglected, but they could be accounted for in calculating the force component Ry.
EXAMPLE 3.8
As shown in Fig. 3.8a, a fixed vane turns a water jet of area A through an angle θ
without changing its velocity magnitude. The flow is steady, pressure is pa everywhere,
and friction on the vane is negligible. (a) Find the components Fx and Fy of the applied
vane force. (b) Find expressions for the force magnitude F and the angle ϕ between F
and the horizontal; plot them versus θ.
Solution
Part (a)
The control volume selected in Fig. 3.8a cuts through the inlet and exit of the jet and through
the vane support, exposing the vane force F. Since there is no cut along the vane–jet
interface, vane ­friction is internally self-canceling. The pressure force is zero in the uniform
atmosphere. We ­neglect the weight of fluid and the vane weight within the control volume.
Then Eq. (3.40) reduces to


Fvane = m2V2 − m1V1
158
Chapter 3 Integral Relations for a Control Volume
y
V
x
pa
2
V
mV
θ
1
CV
Fy
F
ϕ
θ
Fx
Fig. 3.8 Net applied force on a fixed
jet-turning vane: (a) geometry of the
vane turning the water jet; (b) vector
diagram for the net force.
(a)
mV
F
(b)
But the magnitude V1 = V2 = V as given, and conservation of mass for the streamtube



requires m1 = m2 = m = ρAV. The vector diagram for force and momentum change

becomes an isosceles triangle with legs mV and base F, as in Fig. 3.8b. We can readily
find the force components from this diagram:


Fx = mV(cos θ − 1) Fy = mV sin θ
Ans. (a)

2
where m V = ρAV for this case. This is the desired result.
Part (b)
The force magnitude is obtained from part (a):
θ


F = (F2x + F2y ) 1/2 = mV[sin2θ + (cos θ − 1) 2 ] 1/2 = 2mV sin 2
Ans. (b)
2.0
F
mV
F
mV
1.0
180°
ϕ
ϕ
0
45°
90°
θ
E3.8
135°
180°
90°
From the geometry of Fig. 3.8b we obtain
ϕ = 180° − tan −1
Fy
θ
= 90° + Fx
2
Ans. (b)
These can be plotted versus θ as shown in Fig. E3.8. Two special cases are of interest.
First, the maximum force occurs at θ = 180°—that is, when the jet is turned around
and thrown back in the opposite direction with its momentum completely reversed.
3.4 The Linear Momentum Equation 159

This force is 2mV and acts to the left; that is, ϕ = 180°. Second, at very small turning
angles (θ < 10°) we obtain approximately

F ≈ mVθ
ϕ ≈ 90°
The force is linearly proportional to the turning angle and acts nearly normal to
the jet. This is the principle of a lifting vane, or airfoil, which causes a slight
change in the oncoming flow direction and thereby creates a lift force normal to
the basic flow.
One of the differences between Examples 3.7 and 3.8 is the driving force to
push the fluid. In Example 3.7, the water flow is confined in the pipe and elbow,
and is pressure-driven. The pressure in the control volume at the inlet and outlet
is usually higher than the atmospheric pressure. Hence, the pressure forces must
be accounted for in the momentum equation. For an open jet like in Example 3.8,
the motion of the jet stream is due to the initial momentum of the jet flowing
from the nozzle. The water jet is surrounded by atmospheric pressure. Therefore,
there is no contribution of pressure force at the inlet and outlet of the control
volume.
Friction or Drag Force Exerted by Solid Wall on Fluid Flow
Many engineering applications need to find the friction force exerted by pipe wall
on the fluid flow in a pipe or determine the drag force induced by a flat plate
on the flow over it. The overall friction force or drag can be determined using
the linear momentum equation (3.37). Using the differential equation of linear
momentum in Chap. 5 and the boundary theory in Chap. 7 can also help us to
determine local friction along the wall.
EXAMPLE 3.9
In Fig. 3.9 the plate is parallel to the flow. The stream is not a jet but a broad river,
or free stream, of uniform velocity V = U0i. The pressure is assumed uniform, and
so it has no net force on the plate. The only effect is due to boundary shear. The
no-slip condition at the wall brings the fluid there to a halt, and these slowly moving particles retard their neighbors above, so that at the end of the plate there is a
significant retarded shear layer, or boundary layer, of thickness y = δ. The viscous
stresses along the wall can sum to a finite drag force on the plate. These effects
are illustrated in Fig. 3.9. The problem is to make an integral analysis and find the
drag force D in terms of the flow properties ρ, U0, and δ and the plate dimensions
L and b.9
9
The general analysis of such wall shear problems, called boundary-layer theory, is treated
in Sec. 7.3.
160
Chapter 3 Integral Relations for a Control Volume
p = pa
y
Streamline just
outside the
shear-layer region
U0
Fig. 3.9 Control volume analysis of
drag force on a flat plate due to
boundary shear. The control volume
is bounded by sections 1, 2, 3, and 4.
y=δ
2
y=h
Oncoming
stream
parallel
to plate
U0
3
Boundary layer
where shear stress
is significant
1
u(y)
4
0
x
L
Plate of width b
Solution
Like most practical cases, this problem requires a combined mass and momentum balance. A proper selection of control volume is essential, and we select the four-sided
region from 0 to h to δ to L and back to the origin 0, as shown in Fig. 3.9. Had we
chosen to cut across horizontally from left to right along the height y = h, we would
have cut through the shear layer and exposed unknown shear stresses. Instead we follow
the streamline passing through (x, y) = (0, h), which is outside the shear layer and also
has no mass flow across it. The four control volume sides are thus
1.
2.
3.
4.
From (0, 0) to (0, h): a one-dimensional inlet, V · n = −U0.
From (0, h) to (L, δ ): a streamline, no shear, V · n = 0.
From (L, δ) to (L, 0): a two-dimensional outlet, V · n = +u(y).
From (L, 0) to (0, 0): a streamline just above the plate surface, V · n = 0, shear
forces summing to the drag force −Di acting from the plate onto the retarded fluid.
The pressure is uniform, and so there is no net pressure force. Since the flow is assumed
incompressible and steady, Eq. (3.37) applies with no unsteady term and flows only
across sections 1 and 3:
∫
∫
∑ Fx = −D = ρ u(0, y) (V · n) dA + ρ u(L, y) (V · n) dA
1
=ρ
∫
3
h
0
U0 (−U0 )b dy + ρ
∫
δ
u(L, y) [+u(L, y) ]b dy
0
Evaluating the first integral and rearranging give
D = ρU20 bh − ρb
∫
δ
0
u2dy ∣ x=L (1)
This could be considered the answer to the problem, but it is not useful because the
height h is not known with respect to the shear layer thickness δ. This is found by
applying mass conservation, since the control volume forms a streamtube:
ρ
∫
CS
(V · n) dA = 0 = ρ
∫
h
0
(−U0 )b dy + ρ
∫
δ
0
ub dy ∣ x=L
3.4 The Linear Momentum Equation 161
or
U0h =
∫
δ
0
u dy ∣ x=L (2)
after canceling b and ρ and evaluating the first integral. Introduce this value of h into
Eq. (1) for a much cleaner result:
D = ρb
∫
δ
0
u(U0 − u) dy ∣ x=L Ans. (3)
This result was first derived by Theodore von Kármán in 1921.10 It relates the friction
drag on one side of a flat plate to the integral of the momentum deficit ρu(U0 − u)
across the trailing cross section of the flow past the plate. Since U0 − u vanishes as y
increases, the integral has a finite value. Equation (3) is an example of momentum
integral theory for boundary layers, which is treated in Chap. 7.
Nonuniform Hydrostatic Pressure
In hydrodynamic applications, there are problems where the momentum equation
is applied to a control volume for which the pressure is not uniform on the control surface. In this case, the hydrostatic resultant force on a plane or curved
surface should be calculated using the relations we developed in Chap. 2.
EXAMPLE 3.10
The sluice gate in Fig. E3.10a controls flow in open channels. At sections 1 and 2, the
flow is uniform and the pressure is hydrostatic. Neglecting bottom friction and atmospheric pressure, derive a formula for the horizontal force F required to hold the gate.
Express your final formula in terms of the inlet velocity V1, eliminating V2.
A
Sluice
gate, width b
F
h1
V1
h2
V2
E3.10a
Solution
Choose a control volume, Fig. E3.10b, that cuts through known regions (section 1 and
section 2 just above the bottom, and the atmosphere) and that cuts along regions where
unknown information is desired (the gate, with its force F ).
Assume steady incompressible flow with no variation across the width b. The inlet
and outlet mass flows balance:

m = ρV1h1b = ρV2h2b or V2 = V1 (h1∕h2 )
10
The autobiography of this great twentieth-century engineer and teacher [2] is recommended
for its historical and scientific insight.
162
Chapter 3 Integral Relations for a Control Volume
CV
Gage
pressure
F
ρ gh 2
ρ gh 1
E3.10b
τ ≈0
We may use gage pressures for convenience because a uniform atmospheric pressure
causes no force, as shown earlier in Fig. 3.6. With x positive to the right, equate the
net horizontal force to the x-directed momentum change:
ρ
ρ

ΣFx = −Fgate + gh1 (h1b) − gh2 (h2b) = m (V2 − V1 )
2
2

m = ρh1bV1
Solve for Fgate, and eliminate V2 using the mass flow relation. The desired result is:
Fgate =
ρ
h2 2
h1
gbh21[ 1 − ( ) ] − ρh1bV21 ( − 1)
2
h1
h2
Ans.
This is a powerful result from a relatively simple analysis. Later, in Sec. 10.4, we will
be able to calculate the actual flow rate from the water depths and the gate opening
height.
Moving Control Volume
A nondeforming control volume that translates into a straight line at constant
speed is inertial because there is no acceleration. In this case, the linear momentum equation (3.37) is still applicable, except that the velocity V should be
treated as the relative velocity for an inertial, moving, and nondeforming control
volume. Calculation of the relative velocity is a vector operation in kinematics.
However, if the absolute velocity of A and absolute velocity of B are in the same
direction, the relative velocity can be easily calculated using the algebraic
­relation Vc = VA – VB.
EXAMPLE 3.11
A water jet of velocity Vj impinges normal to a flat plate that moves to the right at
velocity Vc, as shown in Fig. 3.10a. Find the force required to keep the plate moving
at constant velocity if the jet density is 1000 kg/m3, the jet area is 3 cm2, and Vj and
Vc are 20 and 15 m/s, respectively. Neglect the weight of the jet and plate, and assume
steady flow with respect to the moving plate with the jet splitting into an equal upward
and downward half-jet.
3.4 The Linear Momentum Equation 163
1 A1 =
p = pa
CS
1
A
2 j
Ry
CS
Vc
Nozzle
Fig. 3.10 Force on a plate moving at
constant velocity: (a) jet striking a
moving plate normally; (b) control
volume fixed relative to the plate.
Rx
Vj – Vc
Vj
Vc
Aj j
2 A2 =
(a)
1
A
2 j
(b)
Solution
The suggested control volume in Fig. 3.10a cuts through the plate support to expose
the desired forces Rx and Ry. This control volume moves at speed Vc and thus is fixed
relative to the plate, as in Fig. 3.10b. We must satisfy both mass and momentum conservation for the assumed steady flow pattern in Fig. 3.10b. There are two outlets and
one inlet, and Eq. (3.30) applies for mass conservation:


mout = min
or
ρ1A1V1 + ρ2A2V2 = ρj Aj (Vj − Vc )
(1)
We assume that the water is incompressible ρ1 = ρ2 = ρj, and we are given that A1 = A2 =
1
2 Aj. Therefore Eq. (1) reduces to
V1 + V2 = 2(Vj − Vc )
(2)
Strictly speaking, this is all that mass conservation tells us. However, from the symmetry of the jet deflection and the neglect of gravity on the fluid trajectory, we conclude
that the two velocities V1 and V2 must be equal, and hence Eq. (2) becomes
V 1 = V2 = Vj − Vc (3)
This equality can also be predicted by Bernoulli’s equation in Sec. 3.5. For the given
numerical values, we have
V1 = V2 = 20 − 15 = 5 m/s
Now we can compute Rx and Ry from the two components of momentum conservation.
Equation (3.40) applies with the unsteady term zero:
∑ Fx = Rx = m 1u1 + m 2u2 − m juj
(4)



where from the mass analysis, m1 = m2 = 12mj = 12ρj Aj (Vj − Vc ). Now check the flow
directions at each section: u1 = u2 = 0, and uj = Vj − Vc = 5 m/s. Thus Eq. (4) becomes

Rx = −mjuj = −[ρj Aj (Vj − Vc ) ] (Vj − Vc )
(5)
164
Chapter 3 Integral Relations for a Control Volume
For the given numerical values we have
Rx = −(1000 kg/m3 ) (0.0003 m2 ) (5 m/s) 2 = −7.5 (kg · m)/s2 = −7.5 N
Ans.
This acts to the left; that is, it requires a restraining force to keep the plate from accelerating to the right due to the continuous impact of the jet. The vertical force is



Fy = Ry = m1υ1 + m2υ2 − mjυj
Check directions again: υ1 = V1, υ2 = −V2, υj = 0. Thus



Ry = m1 (V1 ) + m2 (−V2 ) = 12 mj (V1 − V2 )
(6)
But since we found earlier that V1 = V2, this means that Ry = 0, as we could expect
from the symmetry of the jet deflection.11 Two other results are of interest. First, the
relative velocity at section 1 was found to be 5 m/s up, from Eq. (3). If we convert
this to absolute motion by adding on the control-volume speed Vc = 15 m/s to the right,
we find that the absolute velocity V1 = 15i + 5j m/s, or 15.8 m/s at an angle of 18.4°
upward, as indicated in Fig. 3.10a. Thus the absolute jet speed changes after hitting
the plate. Second, the computed force Rx does not change if we assume the jet deflects
in all radial directions along the plate surface rather than just up and down. Since the
plate is normal to the x axis, there would still be zero outlet x-momentum flow when
Eq. (4) was rewritten for a radial deflection condition.
Momentum Flux Correction Factor
For flow in a duct, the axial velocity is usually nonuniform, as in Example 3.4.

For this case the simple momentum flow calculation e uρ(V · n) dA = mV = ρAV2
is somewhat in error and should be corrected to ζρAV2, where ζ is the dimensionless momentum flow correction factor, ζ ≥ 1.
The factor ζ accounts for the variation of u2 across the duct section. That is,
we compute the exact flow and set it equal to a flow based on average velocity
in the duct:
∫

ρ u2dA = ζm Vav = ζρAV2av
or
ζ=
1
u 2
dA
A ( Vav )
∫
(3.43a)
Values of ζ can be computed based on typical duct velocity profiles similar to
those in Example 3.4. The results are as follows:
Laminar flow:
Turbulent flow:
u = U0 (1 −
r2
4
ζ= 2)
3
R
r m
u ≈ U0 (1 − )
R
ζ=
(3.43b)
1
1
≤m≤
9
5
(1 + m) 2 (2 + m) 2
2(1 + 2m)(2 + 2m)
(3.43c)
11
Symmetry can be a powerful tool if used properly. Try to learn more about the uses and
misuses of symmetry conditions.
3.4 The Linear Momentum Equation 165
The turbulent correction factors have the following range of values:
Turbulent flow:
m
1
5
1
6
1
7
1
8
1
9
ζ
1.037
1.027
1.020
1.016
1.013
These are so close to unity that they are normally neglected. The laminar correction is often important.
To illustrate a typical use of these correction factors, the solution to Example
3.8 for nonuniform velocities at sections 1 and 2 would be modified as
∑ F = m (ζ2V2 − ζ1V1 )
(3.43d )
Note that the basic parameters and vector character of the result are not changed
at all by this correction.
Linear Momentum Tips
The previous examples make it clear that the vector momentum equation is more
difficult to handle than the scalar mass and energy equations. Here are some
momentum tips to remember:
The momentum relation is a vector equation. The forces and the momentum terms are directional and can have three components. A sketch of
these vectors will be indispensable for the analysis.
The momentum flow terms, such as e V(ρV · n)dA, link two different sign
conventions, so special care is needed. First, the vector coefficient V will have
a sign depending on its direction. Second, the mass flow term (ρV · n) will
have a sign (+ , −) depending on whether it is (out, in). For example, in
Fig. 3.8, the x-components of V2 and V1, u2 and u1, are both positive; that is,
they both act to the right. Meanwhile, the mass flow at (2) is positive (out)
and at (1) is negative (in).
The one-dimensional approximation, Eq. (3.40), is glorious, because nonuniform velocity distributions require laborious integration, as in Eq.
(3.11). Thus the momentum flow correction factors ζ are very useful in
avoiding this integration, especially for pipe flow.
The applied forces ΣF act on all the material in the control volume—that
is, the surfaces (pressure and shear stresses), the solid supports that are cut
through, and the weight of the interior masses. Stresses on non-controlsurface parts of the interior are self-canceling and should be ignored.
If the fluid exits subsonically to an atmosphere, the fluid pressure there is
atmospheric.
Where possible, choose inlet and outlet surfaces normal to the flow, so that
pressure is the dominant force and the normal velocity equals the actual
velocity.
Clearly, with that many helpful tips, substantial practice is needed to achieve
momentum skills.
166
Chapter 3 Integral Relations for a Control Volume
Noninertial Reference Frame12
All previous derivations and examples in this section have assumed that the coordinate system is inertial—that is, at rest or moving at constant velocity. In this
case the rate of change of velocity equals the absolute acceleration of the system,
and Newton’s law applies directly in the form of Eqs. (3.2) and (3.35).
In many cases it is convenient to use a noninertial, or accelerating, coordinate
system. An example would be coordinates fixed to a rocket during takeoff. A
second example is any flow on the earth’s surface, which is accelerating relative
to the fixed stars because of the rotation of the earth. Atmospheric and oceanographic flows experience the so-called Coriolis acceleration, outlined next. It is
typically less than 10−5g, where g is the acceleration of gravity, but its accumulated effect over distances of many kilometers can be dominant in geophysical
flows. By contrast, the Coriolis acceleration is negligible in small-scale problems
like pipe or airfoil flows.
Suppose that the fluid flow has velocity V relative to a noninertial xyz coordinate system, as shown in Fig. 3.11. Then dV/dt will represent a noninertial
acceleration that must be added vectorially to a relative acceleration arel to give
the absolute acceleration ai relative to some inertial coordinate system XYZ, as in
Fig. 3.11. Thus
ai =
dV
+ arel
dt
(3.44)
Since Newton’s law applies to the absolute acceleration,
∑ F = ma i = m (
dV
+ arel)
dt
∑ F − marel = m
or
Particle
Vrel = dr
dt
y
r
x
Ω
Noninertial, moving,
rotating coordinates
Y
R
z
X
Fig. 3.11 Geometry of fixed versus
accelerating coordinates.
Inertial
coordinates
Z
12
dV
dt
This section may be omitted without loss of continuity.
(3.45)
3.4 The Linear Momentum Equation 167
Thus Newton’s law in noninertial coordinates xyz is analogous to adding more
“force” terms −marel to account for noninertial effects. In the most general case,
sketched in Fig. 3.11, the term arel contains four parts, three of which account for
the angular velocity Ω(t) of the inertial coordinates. By inspection of Fig. 3.11,
the absolute displacement of a particle is
Si = r + R
(3.46)
Differentiation gives the absolute velocity
dR
Vi = V +
+ Ω × r(3.47)
dt
A second differentiation gives the absolute acceleration:
ai =
dV d2R dΩ
+ 2 +
× r + 2Ω × V + Ω × (Ω × r)
dt
dt
dt
(3.48)
By comparison with Eq. (3.44), we see that the last four terms on the right represent the additional relative acceleration:
1.
2.
3.
4.
d2R/dt2 is the acceleration of the noninertial origin of coordinates xyz.
(dΩ/dt) × r is the angular acceleration effect.
2Ω × V is the Coriolis acceleration.
Ω 3 (Ω × r) is the centripetal acceleration, directed from the particle
normal to the axis of rotation with magnitude Ω2L, where L is the normal
distance to the axis.13
Equation (3.45) differs from Eq. (3.2) only in the added inertial forces on the
left-hand side. Thus the control volume formulation of linear momentum in
noninertial coordinates merely adds inertial terms by integrating the added relative
acceleration over each differential mass in the control volume:
∑F −
where
∫
CV
arel dm =
arel =
d
dt (
∫
CV
Vρ d 𝒱) +
∫
CS
Vρ(Vr . n) dA (3.49)
d2R dΩ
+
× r + 2Ω × V + Ω × (Ω × r)
dt
dt 2
This is the noninertial analog of the inertial form given in Eq. (3.35). To analyze
such problems, one must know the displacement R and angular velocity Ω of the
noninertial coordinates.
If the control volume is fixed in a moving frame, Eq. (3.49) reduces to
∑F −
∫
CV
arel dm =
d
dt (
∫
CV
Vρ d𝒱) +
∫
Vρ(V · n) dA
CS
(3.50)
In other words, the right-hand side reduces to that of Eq. (3.37).
13
A complete discussion of these noninertial coordinate terms is given, for example, in Ref.
4, pp. 49–51.
168
Chapter 3 Integral Relations for a Control Volume
EXAMPLE 3.12
V(t)
V(t)
Accelerating
control volume
m
g
A classic example of an accelerating control volume is a rocket moving straight up, as

in Fig. E3.12. Let the initial mass be M0, and assume a steady exhaust mass flow m
and exhaust velocity Ve relative to the rocket, as shown. If the flow pattern within the
rocket motor is steady and air drag is neglected, derive the differential equation of
vertical rocket motion V(t) and integrate using the initial condition V = 0 at t = 0.
Solution
The appropriate control volume in Fig. E3.12 encloses the rocket, cuts through the exit
jet, and accelerates upward at rocket speed V(t). The z-momentum Eq. (3.49) becomes
∫
∑ Fz − arel dm =
z
Ve
E3.12
Datum
or
−mg − m
d
dt (
dV

= 0 + m (−Ve )
dt
∫

w dm
CV
)

+ (mw) e

with m = m(t) = M0 − mt
The term arel = dV/dt of the rocket. The control volume integral vanishes because of
the steady rocket flow conditions. Separate the variables and integrate, assuming V =
0 at t = 0:

V
t
t
dt
mt

dV = m Ve
− gt Ans.
 − g dt or V(t) = −Veln (1 −
M0 − mt
M0 )
0
0
0
∫
∫
∫
This is a classic approximate formula in rocket dynamics. The first term is positive
and, if the fuel mass burned is a large fraction of initial mass, the final rocket velocity
can exceed Ve.
3.5 Frictionless Flow: The Bernoulli Equation
A classic linear momentum analysis is a relation between pressure, velocity, and
elevation in a frictionless flow, now called the Bernoulli equation. It was stated
(vaguely) in words in 1738 in a textbook by Daniel Bernoulli. A complete derivation of the equation was given in 1755 by Leonhard Euler. The Bernoulli equation
is very famous and very widely used, but one should be wary of its restrictions—
all fluids are viscous and thus all flows have friction to some extent. To use the
Bernoulli equation correctly, one must confine it to regions of the flow that are
nearly frictionless. This section (and, in more detail, Chap. 8) will address the
proper use of the Bernoulli relation.
Consider Fig. 3.12, which is an elemental fixed streamtube control volume
of variable area A(s) and length ds, where s is the streamline direction. The
properties (ρ, V, p) may vary with s and time but are assumed to be uniform
over the cross section A. The streamtube orientation θ is arbitrary, with an
elevation change dz = ds sin θ. Friction on the streamtube walls is shown and
then neglected—a very restrictive assumption. Note that the limit of a vanishingly small area means that the streamtube is equivalent to a streamline of the
flow. Bernoulli’s equation is valid for both and is usually stated as holding
“along a streamline” in frictionless flow.
3.5 Frictionless Flow: The Bernoulli Equation 169
dp
A + dA
p+
ρ + dρ
V + dV
τ=0
A
p + dp
ds
S
0
dz
θ
Fig. 3.12 The Bernoulli equation for
frictionless flow along a streamline:
(a) forces and flows; (b) net pressure
force after uniform subtraction of p.
dp
dp
CV
ρ, V
0
d W ≈ ρg d𝒱
p
dFs ≈ 12 dp dA
(b)
(a)
Conservation of mass [Eq. (3.20)] for this elemental control volume yields
d
dt (
∂ρ



ρ d𝒱) + mout − min = 0 ≈
d𝒱 + dm
∂t
CV
∫

where m = ρAV and d𝒱 ≈ A ds. Then our desired form of mass conservation is
∂ρ

dm = d(ρAV ) = − A ds
∂t
(3.51)
This relation does not require an assumption of frictionless flow.
Now write the linear momentum relation [Eq. (3.37)] in the streamwise direction:
∑dFs =
d
dt (
∂



Vρ d𝒱) + (mV ) out − (mV ) in ≈
(ρV ) A ds + d(mV)
∂t
CV
∫
where Vs = V itself because s is the streamline direction. If we neglect the shear
force on the walls (frictionless flow), the forces are due to pressure and gravity. The
streamwise gravity force is due to the weight component of the fluid within the
control volume:
dFs, grav = −dW sin θ = −γA ds sin θ = −γA dz
The pressure force is more easily visualized, in Fig. 3.12b, by first subtracting a
uniform value p from all surfaces, remembering from Fig. 3.6 that the net force
is not changed. The pressure along the slanted side of the streamtube has a streamwise component that acts not on A itself but on the outer ring of area increase
dA. The net pressure force is thus
dFs,press = 12 dp dA − dp(A + dA) ≈ −A dp
to first order. Substitute these two force terms into the linear momentum relation:
∑ dFs = −γA dz − A dp =
=
∂

(ρV) A ds + d(mV)
∂t
∂ρ
∂V


VA ds +
ρA ds + m dV + V dm
∂t
∂t
170
Chapter 3 Integral Relations for a Control Volume
The first and last terms on the right cancel by virtue of the continuity relation
[Eq. (3.51)]. Divide what remains by ρA and rearrange into the final desired
relation:
dp
∂V
+ V dV + g dz = 0
ds +
ρ
∂t
(3.52)
This is Bernoulli’s equation for unsteady frictionless flow along a streamline. It
is in differential form and can be integrated between any two points 1 and 2 on
the streamline:
∫
2
∂V
ds +
∂t
1
2
∫ dpρ + 12 (V − V ) + g(z
1
2
2
2
1
2
− z1 ) = 0
(3.53)
Steady Incompressible Flow
To evaluate the two remaining integrals, one must estimate the unsteady effect
∂V/∂t and the variation of density with pressure. At this time we consider only
steady (∂V/∂t = 0) incompressible (constant-density) flow, for which Eq. (3.53)
becomes
p2 − p1 1 2
+ (V2 − V21 ) + g(z2 − z1 ) = 0
ρ
2
or
p1 1 2
p2 1 2
+ V1 + gz1 =
+ V2 + gz2 = const ρ
ρ
2
2
(3.54)
This is the Bernoulli equation for steady frictionless incompressible flow along
a streamline.
Bernoulli Interpreted as an Energy Relation
The Bernoulli relation, Eq. (3.54), is a classic momentum result, Newton’s law
for a frictionless, incompressible fluid. It may also be interpreted, however, as an
idealized energy relation. The changes from 1 to 2 in Eq. (3.54) represent reversible pressure work, kinetic energy change, and potential energy change. The fact
that the total remains the same means that there is no energy exchange due to
viscous dissipation, heat transfer, or shaft work. Section 3.7 will add these effects
by making a control volume analysis of the first law of thermodynamics.
Restrictions on the Bernoulli Equation
The Bernoulli equation is a momentum-based force relation and was derived
using the following restrictive assumptions:
1. Steady flow: a common situation, application to most flows in this text.
2. Incompressible flow: appropriate if the flow Mach number is less than 0.3.
This restriction is removed in Chap. 9 by allowing for compressibility.
3. Frictionless flow: restrictive—solid walls and mixing introduce friction effects.
3.5 Frictionless Flow: The Bernoulli Equation 171
4. Flow along a single streamline: different streamlines may have different
“Bernoulli constants” wo = p/ρ + V2/2 + gz, but this is rare. In most cases,
as we shall prove in Chap. 4, a frictionless flow region is irrotational; that
is, curl(V) = 0. For irrotational flow, the Bernoulli constant is the same
everywhere.
The Bernoulli derivation does not account for possible energy exchange due to
heat or work. These thermodynamic effects are accounted for in the steady flow
energy equation. We are thus warned that the Bernoulli equation may be modified
by such an energy exchange.
Figure 3.13 illustrates some practical limitations on the use of Bernoulli’s equation (3.54). For the wind tunnel model test of Fig. 3.13a, the Bernoulli equation is
valid in the core flow of the tunnel but not in the tunnel wall boundary layers, the
model surface boundary layers, or the wake of the model, all of which are regions
with high friction.
In the propeller flow of Fig. 3.13b, Bernoulli’s equation is valid both upstream
and downstream, but with a different constant w0 = p/ρ + V2/2 + gz, caused by
the work addition of the propeller. The Bernoulli relation (3.54) is not valid near
the propeller blades or in the helical vortices (not shown, see Fig. 1.14) shed
downstream of the blade edges. Also, the Bernoulli constants are higher in the
flowing “slipstream” than in the ambient atmosphere because of the slipstream
kinetic energy.
Ambient
air
Valid
Model
Valid,
new
constant
Valid
Valid
Invalid
Invalid
(a)
(b)
Valid, new
constant
Valid
Fig. 3.13 Illustration of regions of
validity and invalidity of the
­Bernoulli equation: (a) tunnel
model, (b) propeller, (c) chimney.
Invalid
(c)
172
Chapter 3 Integral Relations for a Control Volume
For the chimney flow of Fig. 3.13c, Eq. (3.54) is valid before and after the
fire, but with a change in Bernoulli constant that is caused by heat addition. The
Bernoulli equation is not valid within the fire itself or in the chimney wall boundary layers.
Jet Exit Pressure Equals Atmospheric Pressure
When a subsonic jet of liquid or gas exits from a duct into the free atmosphere,
it immediately takes on the pressure of that atmosphere. This is a very important
boundary condition in solving Bernoulli problems, since the pressure at that point
is known. The interior of the free jet will also be atmospheric, except for small
effects due to surface tension and streamline curvature.
Stagnation, Static, and Dynamic Pressures
In many incompressible-flow Bernoulli analyses, elevation changes are negligible.
Thus Eq. (3.54) reduces to a balance between pressure and kinetic energy. We can
write this as
p1 +
1 2
1
ρ V1 = p2 + ρ V22 = po = constant
2
2
The quantity po is the pressure at any point in the frictionless flow where the
velocity is zero. It is called the stagnation pressure and is the highest pressure
possible in the flow, if elevation changes are neglected. The place where zerovelocity occurs is called a stagnation point. For example, on a moving aircraft,
the front nose and the wing leading edges are points of highest pressure. The
pressures p1 and p2 are called static pressures, in the moving fluid. The grouping
(1/2)ρV2 has dimensions of pressure and is called the dynamic pressure. A popular device called a Pitot-static tube (Fig. 6.30) measures (po − p) and then calculates V from the dynamic pressure.
Note, however, that one particular zero-velocity condition, no-slip flow along
a fixed wall, does not result in stagnation pressure. The no-slip condition is a
frictional effect, and the Bernoulli equation does not apply.
Hydraulic and Energy Grade Lines
A useful visual interpretation of Bernoulli’s equation is to sketch two grade lines
of a flow. The energy grade line (EGL) shows the height of the total Bernoulli
constant h0 = z + p/γ + V2/(2g). In frictionless flow with no work or heat transfer [Eq. (3.54)] the EGL has constant height. The hydraulic grade line (HGL)
shows the height corresponding to elevation and pressure head z + p/γ—that is,
the EGL minus the velocity head V2/(2g). The HGL is the height to which liquid
would rise in a piezometer tube (see Prob. 2.11) attached to the flow. In an openchannel flow the HGL is identical to the free surface of the water.
Figure 3.14 illustrates the EGL and HGL for frictionless flow at sections 1
and 2 of a duct. The piezometer tubes measure the static pressure head z + p/γ
and thus outline the HGL. The pitot stagnation-velocity tubes measure the total
head z + p/γ + V2/(2g), which corresponds to the EGL. In this particular case
the EGL is constant, and the HGL rises due to a drop in velocity.
3.5 Frictionless Flow: The Bernoulli Equation 173
Energy grade line
Hydraulic grade line
V12
2g
V22
2g
p2
ρg
p1
ρg
2
w
Flo
Constant
Bernoulli
head
z2
z1
1
Fig. 3.14 Hydraulic and energy grade
lines for frictionless flow in a duct.
Arbitrary datum (z = 0)
In more general flow conditions, the EGL will drop slowly due to friction losses
and will drop sharply due to a substantial loss (a valve or obstruction) or due to
work extraction (to a turbine). The EGL can rise only if there is work addition (as
from a pump or propeller). The HGL generally follows the behavior of the EGL
with respect to losses or work transfer, and it rises and/or falls if the velocity
decreases and/or increases.
As mentioned before, no conversion factors are needed in computations with the
­Bernoulli equation if consistent SI or BG units are used, as the following examples will
show.
In all Bernoulli-type problems in this text, we consistently take point 1
upstream and point 2 downstream.
EXAMPLE 3.13
Find a relation between nozzle discharge velocity V2 and tank free surface height h as
in Fig. E3.13. Assume steady frictionless flow.
Solution
As mentioned, we always choose point 1 upstream and point 2 downstream. Try to
choose points 1 and 2 where maximum information is known or desired. Here we select
point 1 as the tank free surface, where elevation and pressure are known, and point 2
as the nozzle exit, where again pressure and elevation are known. The two unknowns
are V1 and V2.
174
Chapter 3 Integral Relations for a Control Volume
V12
2g
EGL
1
V1
HGL
h = z1 – z2
V2
2
E3.13
Open jet:
p2 = pa
Mass conservation is usually a vital part of Bernoulli analyses. If A1 is the tank
cross section and A2 the nozzle area, this is approximately a one-dimensional flow with
constant density, Eq. (3.30):
A1V1 = A2V2
(1)
Bernoulli’s equation (3.54) gives
p1 1 2
p2 1 2
+ V1 + gz1 =
+ V2 + gz2
ρ 2
ρ 2
But since sections 1 and 2 are both exposed to atmospheric pressure p1 = p2 = pa, the
pressure terms cancel, leaving
V22 − V21 = 2g(z1 − z2 ) = 2gh
(2)
Eliminating V1 between Eqs. (1) and (2), we obtain the desired result:
V22 =
2gh
1 − A22/A21
Ans. (3)
Generally the nozzle area A2 is very much smaller than the tank area A1, so that the
ratio A22 /A21 is doubly negligible, and an accurate approximation for the outlet velocity
is
V2 ≈ (2gh) 1/2
Ans. (4)
This formula, discovered by Evangelista Torricelli in 1644, states that the discharge
velocity equals the speed that a frictionless particle would attain if it fell freely from
point 1 to point 2. In other words, the potential energy of the surface fluid is entirely
converted to kinetic energy of efflux, which is consistent with the neglect of friction
and the fact that no net pressure work is done. Note that Eq. (4) is independent of the
fluid density, a characteristic of gravity-driven flows.
Except for the wall boundary layers, the streamlines from 1 to 2 all behave in
the same way, and we can assume that the Bernoulli constant h0 is the same for all
3.5 Frictionless Flow: The Bernoulli Equation 175
the core flow. However, the outlet flow is likely to be nonuniform, not one-dimensional, so that the average velocity is only approximately equal to Torricelli’s result.
The engineer will then adjust the formula to include a dimensionless discharge
coefficient cd:
(V2 ) av =
Q
= cd (2gh) 1/2
A2
(5)
As discussed in Sec. 6.12, the discharge coefficient of a nozzle varies from about 0.6
to 1.0 as a function of (dimensionless) flow conditions and nozzle shape.
Surface Velocity Condition for a Large Tank
Many Bernoulli, and also steady flow energy, problems involve liquid flow from
a large tank or reservoir, as in Example 3.13. If the outflow is small compared
to the volume of the tank, the surface of the tank hardly moves. Therefore these
problems are analyzed assuming zero velocity at the tank surface. The pressure
at the top of the tank or reservoir is assumed to be atmospheric.
Before proceeding with more examples, we should note carefully that a solution by Bernoulli’s equation (3.54) does not require a second control volume
analysis, only a selection of two points 1 and 2 along a given streamline. The
control volume was used to derive the differential relation (3.52), but the integrated form (3.54) is valid all along the streamline for frictionless flow with no
heat transfer or shaft work, and a control volume is not necessary.
A classic Bernoulli application is the familiar process of siphoning a fluid
from one container to another. No pump is involved; a hydrostatic pressure difference provides the motive force. We analyze this in the following example.
EXAMPLE 3.14
Consider the water siphon shown in Fig. E3.14. Assuming that Bernoulli’s equation is
valid, (a) find an expression for the velocity V2 exiting the siphon tube. (b) If the tube
is 1 cm in diameter and z1 = 60 cm, z2 = −25 cm, z3 = 90 cm, and z4 = 35 cm, e­ stimate
the flow rate in cm3/s.
z3
z1
z4
z =0
E3.14
z2
V2
176
Chapter 3 Integral Relations for a Control Volume
Solution
∙Assumptions: Frictionless, steady, incompressible flow. Write Bernoulli’s equation
starting from where information is known (the surface, z1) and proceeding to where
information is desired (the tube exit, z2).
p2 V22
p1 V21
+
+ gz1 =
+
+ gz2
ρ
ρ
2
2
Note that the velocity is approximately zero at z1, and a streamline goes from z1 to z2.
Note further that p1 and p2 are both atmospheric, p = patm, and therefore cancel. (a)
Solve for the exit velocity from the tube:
V2 = √2g(z1 − z2 )
Ans. (a)
The velocity exiting the siphon increases as the tube exit is lowered below the tank
surface. There is no siphon effect if the exit is at or above the tank surface. Note that
z3 and z4 do not directly enter the analysis. However, z3 should not be too high because
the pressure there will be lower than atmospheric, and the liquid might vaporize. (b)
For the given numerical information, we need only z1 and z2 and calculate, in SI units,
V2 = √2(9.81 m/s2 ) [0.6 m − (−0.25) m] = 4.08 m/s
Q = V2A2 = (4.08 m/s) (π/4) (0.01 m) 2 = 321 E–6 m3/s = 321 cm3/s
Ans. (b)
∙ Comments: Note that this result is independent of the density of the fluid. As an
exercise, you may check that, for water (998 kg/m3), p3 is 11,300 Pa below atmospheric pressure.
In Chap. 6 we will modify this example to include friction effects.
EXAMPLE 3.15
A constriction in a pipe will cause the velocity to rise and the pressure to fall at section 2
in the throat. The pressure difference is a measure of the flow rate through the pipe. The
smoothly necked-down system shown in Fig. E3.15 is called a venturi tube. Find an
­expression for the mass flow in the tube as a function of the pressure change.
p1
1
HGL
p2
2
E3.15
Solution
Bernoulli’s equation is assumed to hold along the center streamline:
p1 1 2
p2
+ 2 V1 + gz1 =
+ 12 V22 + gz2
ρ
ρ
3.5 Frictionless Flow: The Bernoulli Equation 177
If the tube is horizontal, z1 = z2 and we can solve for V2:
V22 − V21 =
2 Δp
ρ
Δp = p1 − p2 (1)
We relate the velocities from the incompressible continuity relation:
A1V1 = A2V2
V1 = β2V2 β =
or
D2
D1
(2)
Combining (1) and (2), we obtain a formula for the velocity in the throat:
V2 = [
ρ(1 − β ) ]
1/2
2 Δp
4
(3)
The mass flow is given by
2ρ Δp 1/2

m = ρA2V2 = A2 (
1 − β4 )
(4)


This is the ideal frictionless mass flow. In practice, we measure mactual = cd mideal and
­correlate the dimensionless discharge coefficient cd.
EXAMPLE 3.16
A 10-cm fire hose with a 3-cm nozzle discharges 1.5 m3/min to the atmosphere. Assuming frictionless flow, find the force FB exerted by the flange bolts to hold the nozzle
on the hose.
Solution
We use Bernoulli’s equation and continuity to find the pressure p1 upstream of the
nozzle, and then we use a control volume momentum analysis to compute the bolt
force, as in Fig. E3.16.
1
2
Water:
1000 kg/m3
pa = 0 (gage)
p1
0
2
1
D2 = 3 cm
D1 = 10 cm
E3.16
0
FB
(a)
CV
x
1
2
FB
0
Control volume
(b)
178
Chapter 3 Integral Relations for a Control Volume
The flow from 1 to 2 is a constriction exactly similar in effect to the venturi in
Example 3.15, for which Eq. (1) gave
p1 = p2 + 12ρ(V22 − V21 )
(1)
3
3
The velocities are found from the known flow rate Q = 1.5 m /min or 0.025 m /s:
Q
0.025 m3/s
V2 =
=
= 35.4 m/s
A2 (π/4) (0.03 m) 2
V1 =
Q
0.025 m3/s
=
= 3.2 m/s
A1 (π/4) (0.1 m) 2
We are given p2 = pa = 0 gage pressure. Then Eq. (1) becomes
p1 = 12 (1000 kg/m3 ) [ (35.42 − 3.22 )m2/s2 ]
= 620,000 kg/(m · s2 ) = 620,000 Pa gage
The control volume force balance is shown in Fig. E3.16b:
∑ Fx = −FB + p1A1
and the zero gage pressure on all other surfaces contributes no force. The x-momentum


flow is +mV2 at the outlet and −mV1 at the inlet. The steady flow momentum relation
(3.40) thus gives

−FB + p1A1 = m (V2 − V1 )

or
FB = p1A1 − m (V2 − V1 )
(2)
Substituting the given numerical values, we find

m = ρQ = (1000 kg/m3 ) (0.025 m3/s) = 25 kg/s
π
π
A1 = D21 = (0.1 m) 2 = 0.00785 m2
4
4
FB = (620,000 N/m2 ) (0.00785 m2 ) − (25 kg/s) [ (35.4 − 3.2)m/s]
= 4872 N − 805 (kg · m)/s2 = 4067 N (915 lbf)
Ans.
Notice from these examples that the solution of a typical problem involving
­ ernoulli’s equation almost always leads to a consideration of the continuity equation
B
as an equal partner in the analysis. The only exception is when the complete velocity
distribution is already known from a previous or given analysis, but that means the
continuity relation has already been used to obtain the given information. The point
is that the continuity relation is always an important element in a flow analysis.
3.6 The Angular Momentum Theorem14
A control volume analysis can be applied to the angular momentum relation, Eq.
(3.3), by letting our dummy variable B be the angular-momentum vector H. However, since the system considered here is typically a group of nonrigid fluid
14
This section may be omitted without loss of continuity.
3.6 The Angular Momentum Theorem 179
particles of variable velocity, the concept of mass moment of inertia is of no help,
and we have to calculate the instantaneous angular momentum by integration over
the elemental masses dm. If O is the point about which moments are desired, the
angular momentum about O is given by
Ho =
∫
syst
(r × V) dm
(3.55)
where r is the position vector from 0 to the elemental mass dm and V is the
velocity of that element. The amount of angular momentum per unit mass is thus
seen to be
β=
dHo
=r×V
dm
The Reynolds transport theorem (3.16) then tells us that
dHo
dt
∣
syst
=
d
dt [
∫
CV
(r × V)ρ d𝒱 ] +
∫
CS
(r × V)ρ(Vr · n) dA
(3.56)
for the most general case of a deformable control volume. But from the angular
momentum theorem (3.3), this must equal the sum of all the moments about point
O applied to the control volume
dHo
=
dt
∑ Mo = ∑ (r × F) o
Note that the total moment equals the summation of moments of all applied forces
about point O. Recall, however, that this law, like Newton’s law (3.2), assumes
that the particle velocity V is relative to an inertial coordinate system. If not, the
moments about point O of the relative acceleration terms arel in Eq. (3.49) must
also be included:
∑ Mo = ∑ (r × F) o −
∫
CV
(r × arel ) dm
(3.57)
where the four terms constituting arel are given in Eq. (3.49). Thus the most
general case of the angular momentum theorem is for a deformable control volume associated with a noninertial coordinate system. We combine Eqs. (3.56) and
(3.57) to obtain
∑ (r × F) o −
∫
CV
(r × arel ) dm =
d
dt [
∫
CV
(r × V)ρ d𝒱 +
∫
(r × V)ρ(Vr · n) dA
CS
For a nondeformable inertial control volume, this reduces to
∑ M0 =
∂
∂t [
∫
CV
(r × V)ρ d𝒱] +
∫
CS
(r × V)ρ(V · n) dA (3.58)
(3.59)
180
Chapter 3 Integral Relations for a Control Volume
Further, if there are only one-dimensional inlets and exits, the angular momentum
flow terms evaluated on the control surface become
∫
CS
(r × V)ρ(V · n)dA =
∑ (r × V) out m out − ∑ (r × V) in m in (3.60)
Although at this stage the angular momentum theorem can be considered a supplementary topic, it has direct application to many important fluid flow problems
involving torques or moments. A particularly important case is the analysis of
rotating fluid flow devices, usually called turbomachines (Chap. 11).
EXAMPLE 3.17
As shown in Fig. E3.17a, a pipe bend is supported at point A and connected to a flow
system by flexible couplings at sections 1 and 2. The fluid is incompressible, and
ambient pressure pa is zero. (a) Find an expression for the torque T that must be resisted
by the support at A, in terms of the flow properties at sections 1 and 2 and the distances
h1 and h2. (b) Compute this torque if D1 = D2 = 3 in, p1 = 100 lbf/in2 gage, p2 = 80
lbf/in2 gage, V1 = 40 ft/s, h1 = 2 in, h2 = 10 in, and ρ = 1.94 slugs/ft3.
A
1
h1
p1, V1, A1
h2
pa = 0
ρ = constant
V2 , A 2 , p 2
2
E3.17a
Solution
Part (a)
The control volume chosen in Fig. E3.17b cuts through sections 1 and 2 and through
the support at A, where the torque TA is desired. The flexible couplings description
specifies that there is no torque at either section 1 or 2, and so the cuts there expose
no moments. For the angular momentum terms r × V, r should be taken from point
A to sections 1 and 2. Note that the gage pressure forces p1A1 and p2A2 both have
moments about A. Equation (3.59) with one-dimensional flow terms becomes
∑ MA = TA + r1 × (−p1A1n1 ) + r2 × (−p2A2n2 )


= (r2 × V2 ) (+mout ) + (r1 × V1 ) (−min )
(1)
3.6 The Angular Momentum Theorem 181
A
r1
V1
p1A1
V2
θ2
TA
r2
h2 = r2 sin θ 2
r2
r2 V2 = h 2 V2
V2
p2 A 2
V1
θ1
h1 = r1 sin θ 1
r1 V1 = h 1 V1
CV
E3.17b
r1
E3.17c
Figure E3.17c shows that all the cross products are associated with either r1 sin θ1 = h1
or r2 sin θ2 = h2, the perpendicular distances from point A to the pipe axes at 1 and 2.


Remember that min = mout from the steady flow continuity relation. In terms of
counterclockwise moments, Eq. (1) then becomes

TA + p1A1h1 − p2A2h2 = m (h2V2 − h1V1 )
(2)
Rewriting this, we find the desired torque to be


TA = h2 ( p2A2 + mV2 ) − h1 ( p1A1 + mV1 )
Ans. (a) (3)
counterclockwise. The quantities p1 and p2 are gage pressures. Note that this result is
independent of the shape of the pipe bend and varies only with the properties at sections 1 and 2 and the distances h1 and h2.15
Part (b)
For the numerical example, convert all data to BG units:
lbf
lbf
lbf
lbf
= 14,400 2 p2 = 80 2 = 11,520 2
in2
ft
in
ft
slug
10
h2 = 10 in =
ft ρ = 1.94 3
12
ft
D1 = D2 = 3 in = 0.25 ft p1 = 100
h1 = 2 in =
2
ft
12
The inlet and exit areas are the same, A1 = A2 = (π/4)(0.25 ft)2 = 0.0491 ft2. Since the
density is constant, we conclude from mass conservation, ρA1V1 = ρA2V2, that V1 = V2
= 40 ft/s. The mass flow is
slug
slug
ft

m = ρ A1V1 = (1.94 3 ) (0.0491 ft2 ) (40 ) = 3.81
s
s
ft
∙ Evaluation of the torque: The data can now be substituted into Eq. (3):
slug
10
lbf
ft
TA = ( ft)[(11,520 2 ) (0.0491 ft2 ) + (3.81
40 )]
)
(
s
s
12
ft
slug
2
lbf
ft
ft
14,400 2 ) (0.0491 ft2 ) + (3.81
40
s ) ( s )]
12 )[(
ft
= 598 ft · lbf − 143 ft · lbf = 455 ft · lbf counterclockwise
Ans. (b)
−(
15
Indirectly, the pipe bend shape probably affects the pressure change from p1 to p2.
182
Chapter 3 Integral Relations for a Control Volume
∙ Comments: The use of standard BG units is crucial when combining dissimilar
terms, such as pressure times area and mass flow times velocity, into proper additive
units for a numerical solution.
EXAMPLE 3.18
Figure 3.15 shows a schematic of a centrifugal pump. The fluid enters axially and passes
through the pump blades, which rotate at angular velocity ω; the velocity of the fluid
is changed from V1 to V2 and its pressure from p1 to p2. (a) Find an expression for the
torque To that must be applied to these blades to maintain this flow. (b) The power supplied to the pump would be P = ωTo. To illustrate numerically, suppose r1 = 0.2 m,
r2 = 0.5 m, and b = 0.15 m. Let the pump rotate at 600 r/min and deliver water at
2.5 m3/s with a density of 1000 kg/m3. Compute the torque and power supplied.
Vn2
Vn1
Blade
Vt2
Vt1
2
z,k
Inflow
1
r
r2
O
ω
r1
Blade
Fig. 3.15 Schematic of a simplified
centrifugal pump.
Pump
blade
shape
CV
Width b
Solution
Part (a)
The control volume is chosen to be the annular region between sections 1 and 2 where
the flow passes through the pump blades (see Fig. 3.15). The flow is steady and
assumed incompressible. The contribution of pressure to the torque about axis O is
zero since the pressure forces at 1 and 2 act radially through O. Equation (3.59)
becomes
∑ Mo = To = (r2 × V2 )m out − (r1 × V1 )m in
(1)
3.6 The Angular Momentum Theorem 183
where steady flow continuity tells us that


min = ρVn12πr1b = mout = ρVn22πr2b = ρQ
The cross product r × V is found to be clockwise about O at both sections:
r2 × V2 = r2Vt2 sin 90° k = r2Vt2 k clockwise
r1 × V1 = r1Vt1k clockwise
Equation (1) thus becomes the desired formula for torque:
To = ρQ (r2Vt2 − r1Vt1 )k clockwise
Ans. (a) (2a)
This relation is called Euler’s turbine formula. In an idealized pump, the inlet and
outlet tangential velocities would match the blade rotational speeds Vt1 = ωr1 and Vt2
= ωr2. Then the formula for torque supplied becomes
To = ρQω (r22 − r21 )
clockwise
(2b)
Part (b)
Convert ω to 600(2π/60) = 62.8 rad/s. The normal velocities are not needed here but
follow from the flow rate
Vn1 =
Q
2.5 m3/s
=
= 13.3 m/s
2πr1b 2π (0.2 m) (0.15 m)
Vn2 =
Q
2.5
=
= 5.3 m/s
2πr2b 2π (0.5) (0.15)
For the idealized inlet and outlet, tangential velocity equals tip speed:
2
Absolute outlet
velocity
V2 = V0i – Rωi
Vt1 = ωr1 = (62.8 rad/s) (0.2 m) = 12.6 m/s
Vt2 = ωr2 = 62.8(0.5) = 31.4 m/s
Equation (2a) predicts the required torque to be
To = (1000 kg/m3 ) (2.5 m3/s) [ (0.5 m) (31.4 m/s) − (0.2 m) (12.6 m/s) ]
= 33,000 (kg · m2 )/s2 = 33,000 N · m
Ans.
The power required is
R
y
ω
P = ωTo = (62.8 rad/s) (33,000 N · m) = 2,070,000 (N · m)/s
= 2.07 MW (2780 hp)
CV
Retarding
torque T0
Ans.
In actual practice the tangential velocities are considerably less than the impeller-tip
speeds, and the design power requirements for this pump may be only 1 MW or less.
x
O
Inlet velocity
Q
k
V0 =
A pipe
Fig. 3.16 View from above of a single arm of a rotating lawn sprinkler.
EXAMPLE 3.19
Figure 3.16 shows a lawn sprinkler arm viewed from above. The arm rotates about O at
constant angular velocity ω. The volume flow entering the arm at O is Q, and the fluid
is incompressible. There is a retarding torque at O, due to bearing friction, of amount
−Tok. Find an expression for the rotation ω in terms of the arm and flow properties.
184
Chapter 3 Integral Relations for a Control Volume
Solution
The entering velocity is V0k, where V0 = Q/Apipe. Equation (3.59) applies to the control
volume sketched in Fig. 3.16 only if V is the absolute velocity relative to an inertial
frame. Thus the exit velocity at section 2 is
V2 = V0i − Rωi
Equation (3.59) then predicts that, for steady flow,
∑ Mo = −Tok = (r2 × V2 )m out − (r1 × V1 )m in
(1)


where, from continuity, mout = min = ρQ. The cross products with reference to point O
are
r2 × V2 = Rj × (V0 − Rω)i = (R2ω − RV0 )k
r1 × V1 = 0j × V0k = 0
Equation (1) thus becomes
−Tok = ρQ(R2ω − RV0 )k
ω=
Vo
To
−
R
ρQR2
Ans.
The result may surprise you: Even if the retarding torque To is negligible, the arm
rotational speed is limited to the value V0/R imposed by the outlet speed and the arm
length.
3.7 The Energy Equation16
As our fourth and final basic law, we apply the Reynolds transport theorem (3.12)
to the first law of thermodynamics, Eq. (3.5). The dummy variable B becomes
energy E, and the energy per unit mass is β = dE/dm = e. Equation (3.5) can
then be written for a fixed control volume as follows:17
dQ dW dE
d
−
=
= (
dt
dt
dt
dt
∫
CV
eρ d𝒱 ) +
∫
CS
eρ(V · n) dA
(3.61)
Recall that positive Q denotes heat added to the system and positive W denotes
work done by the system.
The system energy per unit mass e may be of several types:
e = einternal + ekinetic + epotential
We consider the three terms as discussed in Eq. (1.9), with z defined as “up”:
e = û + 12V2 + gz
(3.62)
16
This section should be read for information and enrichment even if you lack formal background
in thermodynamics.
17
The energy equation for a deformable control volume is rather complicated and is not discussed
here. See Refs. 4 and 5 for further details.
3.7 The Energy Equation 185
The heat and work terms could be examined in detail. If this were a heat
transfer book, dQ/dt would be broken down into conduction, convection, and
radiation effects and whole chapters written on each (see, for example, Ref. 3).
Here we leave the term untouched and consider it only occasionally.
Using for convenience the overdot to denote the time derivative, we divide the
work term into three parts:







W = Wshaft + Wpress + Wviscous stresses = Ws + Wp + Wυ
The work of gravitational forces has already been included as potential energy in Eq.
(3.62). Other types of work, such as those due to electromagnetic forces, are excluded
here.
The shaft work isolates the portion of the work that is deliberately done by a
machine (pump impeller, fan blade, piston, or the like) protruding through  the
control surface into the control volume. No further specification other than Ws is
desired at this point, but calculations of the work done by turbomachines will be
performed in Chap. 11.
The rate of work Wp done by pressure forces occurs at the surface only; all
work on internal portions of the material in the control volume is by equal and
opposite forces and is self-canceling. The pressure work equals the pressure force
on a small surface element dA times the normal velocity component into the
control volume:

dWp = −( p dA)Vn, in = −p(−V · n) dA
The total pressure work is the integral over the control surface:

Wp =
∫
p(V · n) dA
CS
(3.63)
A cautionary remark: If part of the control surface is the surface of a machine part,
we prefer
to delegate that portion of the pressure to the shaft work term Ws, not

to Wp, which is primarily meant to isolate the fluid flow pressure work terms.
Finally, the shear work due to viscous stresses occurs at the control surface
and consists of the product of each viscous stress (one normal and two tangential)
and the respective velocity component:

dWυ = −τ · V dA

or
Wυ = − τ · V dA
(3.64)
∫
CS
where τ is the stress vector on the elemental surface dA. This term may vanish
or be negligible according to the particular type of surface at that part of the
control volume:
Solid surface. For all parts of the control surface that are solid
 confining
walls, V = 0 from the viscous no-slip condition; hence Wυ = zero
identically.
Surface of a machine. Here the viscous work is contributed by the machine,
and so we absorb this work in the term Ws.
186
Chapter 3 Integral Relations for a Control Volume
An inlet or outlet. At an inlet or outlet, the flow is approximately normal to
the element dA; hence the only viscous work term comes from the normal stress τnnVn dA. Since viscous normal stresses are extremely small in
all but rare cases, such as the interior of a shock wave, it is customary to
neglect viscous work at inlets and outlets of the control volume.
Streamline surface. If the control surface is a streamline, such as the upper
curve in the boundary-layer analysis of Fig. 3.11, the viscous work term
must be evaluated and retained if shear stresses are significant along this
line. In the particular case of Fig. 3.11, the streamline is outside the
boundary layer, and viscous work is negligible.
The net result of this discussion is that the rate-of-work term in Eq. (3.61)
consists essentially of


W = Ws +
∫
p(V · n) dA −
CS
∫
CS
(τ · V) ss dA
(3.65)
where the subscript SS stands for stream surface. When we introduce (3.65) and
(3.62) into (3.61), we find that the pressure work term can be combined with the
energy flow term since both involve surface integrals of V · n. The control volume
energy equation thus becomes



p
∂
Q − Ws − Wυ = (
eρ d 𝒱) +
(e + ρ )ρ(V · n) dA (3.66)
∂t
∫
CV
∫
CS
Using e from (3.62), we see that the enthalpy ĥ = û + p/ρ occurs in the control
surface integral. The final general form for the energy equation for a fixed control
volume becomes



∂
Q − Ws − Wυ =
∂t [
∫
(
CV
û + 12V2 + gz ρd𝒱 +
)
]
∫
CS
(
ĥ + 12V2 + gz ρ(V · n)dA
)

As mentioned, the shear work term Wυ is rarely important.
(3.67)
One-Dimensional Energy-Flux Terms
If the control volume has a series of one-dimensional inlets and outlets, as in
Fig. 3.5, the surface integral in (3.67) reduces to a summation of outlet flows
minus inlet flows:
∫
CS
(ĥ + 12V 2 + gz)ρ(V · n) dA
=
∑ (ĥ + 12V2 + gz) outm out − ∑ (ĥ + 12V2 + gz) inm in (3.68)
where the values of ĥ, 12 V2, and gz are taken to be averages over each cross section.
3.7 The Energy Equation 187
Q=?
150 hp
A steady flow machine (Fig. E3.20) takes in air at section 1 and discharges it at sections 2 and 3. The properties at each section are as follows:
(2)
(1)
(3)
CV
E3.20
EXAMPLE 3.20
Section
A, ft2
Q, ft3/s
1
0.4
100
70
20
1.0
2
1.0
40
100
30
4.0
3
0.25
50
200
?
1.5
T, °F
p, lbf/in2 abs
z, ft
Work is provided to the machine at the rate of 150 hp. Find the pressure p3 in lbf/in2
absolute and the heat transfer Q in Btu/s. Assume that air is a perfect gas with R =
1716 and cp = 6003 ft-lbf/(slug · °R).
Solution
∙ System sketch: Figure E3.20 shows inlet 1 (negative flow) and outlets 2 and 3 (positive
flows).
∙ Assumptions: Steady flow, one-dimensional inlets and outlets, ideal gas, negligible shear
work. The flow is not incompressible. Note that Q1 ≠ Q2 + Q3 because the densities are
different.
∙ Approach: Evaluate the velocities and densities and enthalpies and substitute into Eq.
(3.67). Use BG units for all properties, including the pressures. With Qi given, we evaluate Vi = Qi/Ai:
V1 =
Q1 100 ft3/s
ft
=
= 250
2
s
A1
0.4 ft
V2 =
40 ft3/s
ft
= 40
2
s
1.0 ft
V3 =
50 ft3/s
ft
= 200
2
s
0.25 ft
The densities at sections 1 and 2 follow from the ideal-gas law:
ρ1 =
slug
p1
(20 × 144) lbf/ft2
=
= 0.00317 3
RT1 [1716 ft-lbf/(slug°R) ] [ (70 + 460)°R]
ft
ρ2 =
slug
(30 × 144)
= 0.00450 3
(1716) (100 + 460)
ft
However, p3 is unknown, so how do we find ρ3? Use the steady flow continuity relation:



m1 = m2 + m3 or ρ1Q1 = ρ2Q2 + ρ3Q3 (1)
3
slug
slug
ft
(0.00317 ft3 ) (100 s ) = 0.00450(40) + ρ3 (50) solve for ρ3 = 0.00274 ft3
Knowing ρ3 enables us to find p3 from the ideal-gas law:
slug
ft-lbf
lbf
lbf
p3 = ρ3RT3 = (0.00274 3 ) (1716
(200 + 460°R) = 3100 2 = 21.5 2 Ans.
)
slug
°R
ft
in
ft
∙ Final solution steps: For an ideal gas, simply approximate enthalpies as hi = cpTi.
The shaft work is negative (into the control volume) and viscous work is neglected
for this solid-wall machine:


ft-lbf
ft-lbf
Wυ ≈ 0 Ws = (−150 hp) (550
= −82,500
(work on the system)
s
s-hp )
188
Chapter 3 Integral Relations for a Control Volume
For steady flow, the volume integral in Eq. (3.67) vanishes, and the energy equation
becomes





Q − Ws = −m1 (cpT1 + 12 V21 + gz1 ) + m2 (cpT2 + 12 V22 + gz2 ) + m3 (cpT3 + 12 V23 + gz3 ) (2)
From our continuity calculations in Eq. (1) above, the mass flows are
slug
slug


m1 = ρ1Q1 = (0.00317) (100) = 0.317
m2 = ρ2Q2 = 0.180
s
s
slug

m3 = ρ3Q3 = 0.137
s
It is instructive to separate the flow terms in the energy equation (2) for examination:



Enthalpy flow = cp (−m1T1 + m2T2 + m3T3 )
= (6003) [ (−0.317) (530) + (0.180) (560) + (0.137) (660) ]
= −1,009,000 + 605,000 + 543,000 ≈ +139,000 ft-lbf/s



Kinetic energy flow = 12 (−m1V21 + m2V22 + m3V23 )
= 12 [−0.317(250) 2 + (0.180) (40) 2 + (0.137) (200) 2 ]
= −9900 + 140 + 2740 ≈ −7000 ft-lbf/s



Potential energy flow = g(−m1z1 + m2z2 + m3z3 )
= (32.2) [−0.317(1.0) + 0.180(4.0) + 0.137(1.5) ]
= −10 + 23 + 7 ≈ +20 ft-lbf/s
Equation (2) may now be evaluated for the heat transfer:

Q − (−82,500) = 139,000 − 7,000 + 20

ft-lbf
1 Btu
Btu
or
Q ≈ (+ 49,520
= + 64
s ) ( 778.2 ft-lbf )
s
Ans.
∙ Comments: The heat transfer is positive, which means into the control volume. It is
typical of gas flows that potential energy flow is negligible, enthalpy flow is dominant,
and kinetic energy flow is small unless the velocities are very high (that is, high subsonic or supersonic).
The Steady Flow Energy Equation
For steady flow with one inlet and one outlet, both assumed one-dimensional,
Eq. (3.67) reduces to a celebrated relation used in many engineering analyses.
Let section 1 be the inlet and section 2 the outlet. Then





Q − Ws − Wυ = −m1 (ĥ1 + 12 V21 + gz1 ) + m2 (ĥ2 + 12 V22 + gz2 )
(3.69)



But, from continuity, m1 = m2 = m, we can rearrange (3.69) as follows:
ĥ1 + 12 V21 + gz1 = (ĥ2 + 12 V22 + gz2 ) − q + ws + wυ
(3.70)
 
where q = Q/m
 = dQ/dm, the heat transferred
  to the fluid per unit mass. Similarly, ws = Ws /m = dWs/dm and wυ = Wυ /m = dWυ /dm. Equation (3.70) is a
general form of the steady flow energy equation, which states that the upstream
3.7 The Energy Equation 189
stagnation enthalpy H1 = (h + 12V2 + gz) 1 differs from the downstream value H2
only if there is heat transfer, shaft work, or viscous work as the fluid passes
between sections 1 and 2. Recall that q is positive if heat is added to the control
volume and that ws and wυ are positive if work is done by the fluid on the surroundings.
Each term in Eq. (3.70) has the dimensions of energy per unit mass, or velocity
squared, which is a form commonly used by mechanical engineers. If we divide
through by g, each term becomes a length, or head, which is a form preferred by
civil engineers. The traditional symbol for head is h, which we do not wish to
confuse with enthalpy. Therefore we use internal energy in rewriting the head form
of the energy relation:
p1 û1 V21
p2 û2 V22
+ z1 =
+ z2 − hq + hs + hυ
+
+
+
+
γ
g
γ
g
2g
2g
(3.71)
where hq = q/g, hs = ws /g, and hυ = wv /g are the head forms of the heat added,
shaft work done, and viscous work done, respectively. The term p/γ is called
pressure head, and the term V2/2g is denoted as velocity head.
Friction and Shaft Work in Low-Speed Flow
A common application of the steady flow energy equation is for low-speed
(incompressible) flow through a pipe or duct. A pump or turbine may be included
in the pipe system. The pipe and machine walls are solid, so the viscous work is
zero. Equation (3.71) may be written as
û2 − û1 − q
p1 V21
p2 V22
+
z
=
+ z2 ) +
+
+
1
(γ
)
(
γ
g
2g
2g
(3.72)
Every term in this equation is a length, or head. The terms in parentheses are the
upstream (1) and downstream (2) values of the useful or available head or total head
of the flow, denoted by h0. The last term on the right is the difference (h01 − h02),
which can include pump head input, turbine head extraction, and the friction head
loss hf, always positive. Thus, in incompressible flow with one inlet and one outlet,
we may write
p V2
p V2
+
z
=
+
( γ 2g
)in ( γ + 2g + z)out + hfriction − hpump + hturbine (3.73)
Most of our internal flow problems will be solved with the aid of Eq. (3.73). The
h terms are all positive; that is, friction loss is always positive in real (viscous)
flows, a pump adds energy (increases the left-hand side), and a turbine extracts
energy from the flow. If hp and/or ht are included, the pump and/or turbine must
lie between points 1 and 2. In Chaps. 5 and 6 we shall develop methods of correlating hf losses with flow parameters in pipes, valves, fittings, and other internal flow devices.
190
Chapter 3 Integral Relations for a Control Volume
EXAMPLE 3.21
Gasoline at 20°C is pumped through a smooth 12-cm-diameter pipe 10 km long, at a
flow rate of 75 m3/h (330 gal/min). The inlet is fed by a pump at an absolute pressure
of 24 atm. The exit is at standard atmospheric pressure and is 150 m higher. Estimate
the frictional head loss hf, and compare it to the velocity head of the flow V2/(2g).
(These numbers are quite realistic for liquid flow through long pipelines.)
Solution
∙ Property values: From Table A.3 for gasoline at 20°C, ρ = 680 kg/m3, or γ =
(680)(9.81) = 6670 N/m3.
∙ Assumptions: Steady flow. No shaft work, thus hp = ht = 0. If z1 = 0, then z2 = 150 m.
∙ Approach: Find the velocity and the velocity head. These are needed for comparison.
Then evaluate the friction loss from Eq. (3.73).
∙ Solution steps: Since the pipe diameter is constant, the average velocity is the same
everywhere:
Vin = Vout =
Q
Q
(75 m3/h)/(3600 s/h)
m
=
=
≈ 1.84
2
s
A (π/4)D
(π/4) (0.12 m) 2
Velocity head =
(1.84 m/s) 2
V2
=
≈ 0.173 m
2g 2(9.81 m/s2 )
Substitute into Eq. (3.73) and solve for the friction head loss. Use pascals for the
pressures and note that the velocity heads cancel because of the constant-area pipe.
pin V2in
pout V2out
+
+ zin =
+
+ zout + hf
γ
γ
2g
2g
(24) (101,350 N/m2 )
3
6670 N/m
or
+ 0.173 m + 0 m =
101,350 N/m2
+ 0.173 m + 150 m + hf
6670 N/m3
hf = 364.7 − 15.2 − 150 ≈ 199 m
Ans.
The friction head is larger than the elevation change Δz, and the pump must drive the
flow against both changes, hence the high inlet pressure. The ratio of friction to velocity
head is
hf
199 m
≈
≈ 1150
Ans.
V2/(2g) 0.173 m
∙ Comments: This high ratio is typical of long pipelines. (Note that we did not make direct
use of the 10,000-m pipe length, whose effect is hidden within hf.) In Chap. 6 we can
state this problem in a more direct fashion: Given the flow rate, fluid, and pipe size, what
inlet pres­sure is needed? Our correlations for hf will lead to the estimate pinlet ≈ 24 atm,
as stated here.
EXAMPLE 3.22
Air [R = 1716, cp = 6003 ft · lbf/(slug · °R)] flows steadily, as shown in Fig. E3.22,
through a turbine that produces 700 hp. For the inlet and exit conditions shown,
­estimate (a) the exit velocity V2 and (b) the heat transferred Q in Btu/h.
3.7 The Energy Equation 191
W˙ s = 700 hp
2
1
Turbomachine
D1 = 6 in
p1 = 150
D2 = 6 in
lb/in2
T1 = 300° F
Q˙ ?
p2 = 40 lb/in2
T2 = 35° F
V1 = 100 ft/s
E3.22
Solution
Part (a)
The inlet and exit densities can be computed from the perfect-gas law:
ρ1 =
p1
150(144)
=
= 0.0166 slug/ft3
RT1 1716(460 + 300)
ρ2 =
p2
40(144)
=
= 0.00679 slug/ft3
RT2 1716(460 + 35)
The mass flow is determined by the inlet conditions
π 6 2

m = ρ1A1V1 = (0.0166) ( ) (100) = 0.325 slug/s
4 12
Knowing mass flow, we compute the exit velocity
π 6 2

m = 0.325 = ρ2A2V2 = (0.00679) ( ) V2
4 12
or
V2 = 244 ft/s
Ans. (a)
Part (b)

The steady flow energy equation (3.69) applies with Wυ = 0, z1 = z2, and ĥ = cpT:



Q − Ws = m (cpT2 + 12 V22 − cpT1 − 12 V21 )
Convert the turbine work to foot-pounds-force
per second with the conversion factor

1 hp = 550 ft · lbf/s. The turbine work Ws is positive

Q − 700(550) = 0.325[6003(495) + 12 (244) 2 − 6003(760) − 21 (100) 2 ]
or
= −510,000 ft · lbf/s

Q = −125,000 ft · lbf/s
192
Chapter 3 Integral Relations for a Control Volume
Convert this to British thermal units as follows:

Q = (−125,000 ft · lbf/s)
3600 s/h
778.2 ft · lbf/Btu
= −578,000 Btu/h
Ans. (b)
The negative sign indicates that this heat transfer is a loss from the control volume.
Kinetic Energy Correction Factor
Often the flow entering or leaving a port is not strictly one-dimensional. In particular, the velocity may vary over the cross section, as in Fig. E3.4. In this case
the kinetic energy term in Eq. (3.68) for a given port should be modified by a
dimensionless correction factor α so that the integral can be proportional to the
square of the average velocity through the port:
∫
port
where
Vav =

2
2
( 12 V )ρ(V · n) dA ≡ α( 12 Vav )m
∫
1
u dA
A
for incompressible flow
If the density is also variable, the integration is very cumbersome; we shall not
treat this complication. By letting u be the velocity normal to the port, the first
equation above becomes, for incompressible flow,
1
2
or
∫
ρ u3dA = 12 ραV3avA
α=
1
u 3
dA
A ( Vav )
∫
(3.74)
The term α is the kinetic energy correction factor, having a value of about 2.0
for fully developed laminar pipe flow and from 1.04 to 1.11 for turbulent pipe
flow. The complete incompressible steady flow energy equation (3.73), including
pumps, turbines, and losses, would generalize to
p
p
α 2
α 2
+
+
V
+
z
=
( ρg 2g
)in ( ρg 2g V + z)out + hturbine − hpump + hfriction (3.75)
where the head terms on the right (ht, hp, hf) are all numerically positive. All
additive terms in Eq. (3.75) have dimensions of length {L}. In problems involving turbulent pipe flow, it is common to assume that α ≈ 1.0. To compute numerical values, we can use these approximations to be discussed in Chap. 6:
Laminar flow:
r 2
u = U0 [ 1 − ( ) ]
R
3.7 The Energy Equation 193
from which
Vav = 0.5U0
and
α = 2.0
u ≈ U0 (1 −
Turbulent flow:
(3.76)
r m
R)
m≈
1
7
from which, in Example 3.4,
Vav =
2U0
(1 + m)(2 + m)
Substituting into Eq. (3.74) gives
α=
(1 + m) 3 (2 + m) 3
4 (1 + 3m)(2 + 3m)
(3.77)
and numerical values are as follows:
Turbulent flow:
m
1
5
1
6
1
7
1
8
1
9
α
1.106
1.077
1.058
1.046
1.037
These values are only slightly different from unity and are often neglected in
elementary turbulent flow analyses. However, α should never be neglected in
laminar flow.
EXAMPLE 3.23
A hydroelectric power plant (Fig. E3.23) takes in 30 m3/s of water through its turbine
and discharges it to the atmosphere at V2 = 2 m/s. The head loss in the turbine and
penstock system is hf = 20 m. Assuming turbulent flow, α ≈ 1.06, estimate the power
in MW extracted by the turbine.
1
z1 = 100 m
Water
30 m3/s
z2 = 0 m
2 m/s
E3.23
Turbine
194
Chapter 3 Integral Relations for a Control Volume
Solution
We neglect viscous work and heat transfer and take section 1 at the reservoir surface
(Fig. E3.23), where V1 ≈ 0, p1 = patm, and z1 = 100 m. Section 2 is at the turbine
outlet.
The steady flow energy equation (3.75) becomes, in head form,
p2 α2V22
p1 α1V21
+ z1 =
+ z2 + ht + hf
+
+
γ
γ
2g
2g
pa 1.06(0) 2
pa 1.06(2.0 m/s) 2
+ 100 m =
+ 0 m + ht + 20 m
+
+
γ
γ
2(9.81)
2(9.81 m/s2 )
The pressure terms cancel, and we may solve for the turbine head (which is
positive):
ht = 100 − 20 − 0.2 ≈ 79.8 m
The turbine extracts about 79.8 percent of the 100-m head available from the dam. The
total power extracted may be evaluated from the water mass flow:

P = mws = (ρQ) (ght ) = (998 kg/m3 ) (30 m3/s) (9.81 m/s2 ) (79.8 m)
= 23.4 E6 kg · m2/s3 = 23.4 E6 N · m/s = 23.4 MW
Ans.
The turbine drives an electric generator that probably has losses of about 15 percent,
so the net power generated by this hydroelectric plant is about 20 MW.
EXAMPLE 3.24
The pump in Fig. E3.24 delivers water (62.4 lbf/ft3) at 1.5 ft3/s to a machine at section
2, which is 20 ft higher than the reservoir surface. The losses between 1 and 2 are
given by hf = KV22/(2g), where K ≈ 7.5 is a dimensionless loss coefficient (see Sec. 6.7).
Take α ≈ 1.07. Find the horsepower required for the pump if it is 80 percent efficient.
Machine
p1 = 14.7 lbf/in2 abs
1
Water
E3.24
2
z1 = 0
D2 = 3 in
z2 = 20 ft
p2 = 10 lbf/in2
Pump
hs (negative)
Solution
∙ System sketch: Figure E3.24 shows the proper selection for sections 1 and 2.
∙ Assumptions: Steady flow, negligible viscous work, large reservoir (V1 ≈ 0).
Summary 195
∙ Approach: First find the velocity V2 at the exit, then apply the steady flow energy
equation.
∙ Solution steps: Use BG units, p1 = 14.7(144) = 2117 lbf/ft2 and p2 = 10(144) =
1440 lbf/ft2.
Find V2 from the known flow rate and the pipe diameter:
V2 =
Q
1.5 ft3/s
=
= 30.6 ft/s
A2 (π/4) (3/12 ft) 2
The steady flow energy equation (3.75), with a pump (no turbine) plus z1 ≈ 0 and
V1 ≈ 0, becomes
p2 α2V22
p1 α1V21
V22
+ z1 =
+ z2 − hp + hf , hf = K
+
+
γ
γ
2g
2g
2g
or
p2 − p 1
V22
+ z2 + (α2 + K)
γ
2g
hp =
∙ Comment: The pump must balance four different effects: the pressure change, the
elevation change, the exit jet kinetic energy, and the friction losses.
∙ Final solution: For the given data, we can evaluate the required pump head:
hp =
(1440 − 2117) lbf/ft2
62.4 lbf/ft3
+ 20 + (1.07 + 7.5)
(30.6 ft/s) 2
2(32.2 ft/s2 )
= −11 + 20 + 124 = 133 ft
With the pump head known, the delivered pump power is computed similar to the
turbine in Example 3.23:
lbf
ft3

Ppump = mws = γQhp = (62.4 3 ) (1.5 ) (133 ft)
s
ft
= 12,450
12,450 ft-lbf/s
ft − lbf
=
= 22.6 hp
s
550 ft-lbf/(s−hp)
If the pump is 80 percent efficient, then we divide by the efficiency to find the input
power required:
Pinput =
Ppump
efficiency
=
22.6 hp
= 28.3 hp
0.80
Ans.
∙ Comment: The inclusion of the kinetic energy correction factor α in this case made
a difference of about 1 percent in the result. The friction loss, not the exit jet, was
the dominant parameter.
Summary
This chapter has analyzed the four basic equations of fluid mechanics: conservation of (1) mass, (2) linear momentum, (3) angular momentum, and (4) energy.
The equations were attacked “in the large”—that is, applied to whole regions of
196
Chapter 3 Integral Relations for a Control Volume
a flow. As such, the typical analysis will involve an approximation of the flow
field within the region, giving somewhat crude but always instructive quantitative
results. However, the basic control volume relations are rigorous and correct and
will give exact results if applied to the exact flow field.
There are two main points to a control volume analysis. The first is the selection of a proper, clever, workable control volume. There is no substitute for
experience, but the following guidelines apply. The control volume should cut
through the place where the information or solution is desired. It should cut
through places where maximum information is already known. If the momentum
equation is to be used, it should not cut through solid walls unless absolutely
necessary, since this will expose possible unknown stresses and forces and
moments that make the solution for the desired force difficult or impossible.
Finally, every attempt should be made to place the control volume in a frame of
reference where the flow is steady or quasi-steady, since the steady formulation
is much simpler to evaluate.
The second main point to a control volume analysis is the reduction of the
analysis to a case that applies to the problem at hand. The 24 examples in this
chapter give only an introduction to the search for appropriate simplifying
assumptions. You will need to solve 24 or 124 more examples to become truly
experienced in simplifying the problem just enough and no more. In the meantime, it would be wise for the beginner to adopt a very general form of the
control volume conservation laws and then make a series of simplifications to
achieve the final analysis. Starting with the general form, one can ask a series of
questions:
1.
2.
3.
4.
5.
6.
7.
8.
Is the control volume nondeforming or nonaccelerating?
Is the flow field steady? Can we change to a steady flow frame?
Can friction be neglected?
Is the fluid incompressible? If not, is the perfect-gas law applicable?
Are gravity or other body forces negligible?
Is there heat transfer, shaft work, or viscous work?
Are the inlet and outlet flows approximately one-dimensional?
Is atmospheric pressure important to the analysis? Is the pressure hydrostatic
on any portions of the control surface?
9. Are there reservoir conditions that change so slowly that the velocity and
time rates of change can be neglected?
In this way, by approving or rejecting a list of basic simplifications like these,
one can avoid pulling Bernoulli’s equation off the shelf when it does not apply.
Problems
Most of the problems herein are fairly straightforward. More
difficult or open-ended assignments are labeled with an asterisk. Problems labeled with a computer icon
may require
the use of a computer. The standard end-of-chapter problems
P3.1 to P3.184 (categorized in the problem list here) are followed by
word problems W3.1 to W3.7, fundamentals of engineering (FE)
exam problems FE3.1 to FE3.10, comprehensive problems C3.1 to
C3.5, and design project D3.1.
Problems 197
from a hole in the bottom of area Ao. Use the Reynolds
transport theorem to find an expression for the instantaneous depth change dh/dt.
A spherical tank, of diameter 35 cm, is leaking air
through a 5-mm-diameter hole in its side. The air exits
the hole at 360 m/s and a density of 2.5 kg/m3. Assuming
uniform mixing, (a) find a formula for the rate of change
of average density in the tank and (b) calculate a numerical value for (dρ/dt) in the tank for the given data.
Three pipes steadily deliver water at 20°C to a large exit
pipe in Fig. P3.8. The velocity V2 = 5 m/s, and the exit
flow rate Q4 = 120 m3/h. Find (a) V1, (b) V3, and (c) V4 if
it is known that increasing Q3 by 20 percent would increase Q4 by 10 percent.
Problem Distribution
Section
3.1
3.2
3.3
3.4
3.5
3.6
3.7
Topic
Problems
Basic physical laws; volume flow
The Reynolds transport theorem
Conservation of mass
The linear momentum equation
The Bernoulli equation
The angular momentum theorem
The energy equation
P3.1–P3.5
P3.6–P3.9
P3.10–P3.38
P3.39–P3.109
P3.110–P3.148
P3.149–P3.164
P3.165–P3.184
P3.7
P3.8
Basic physical laws; volume flow
P3.1
Discuss Newton’s second law (the linear momentum
­relation) in these three forms:
∑ F = ma ∑ F =
∑F =
P3.2
d
dt (
∫
system
Vρ d 𝒱)
D2 = 5 cm
Are they all equally valid? Are they equivalent? Are
some forms better for fluid mechanics as opposed to
solid ­mechanics?
Consider the angular momentum relation in the form
∑ MO =
d
dt [
∫
system
(r × V)ρ d 𝒱 ]
What does r mean in this relation? Is this relation valid in
both solid and fluid mechanics? Is it related to the linear
momentum equation (Prob. 3.1)? In what manner?
P3.3 For steady low-Reynolds-number (laminar) flow through
a long tube (see Prob. 1.12), the axial velocity distribution is given by u = C(R2 − r2), where R is the tube radius
and r ≤ R. Integrate u(r) to find the total volume flow Q
through the tube.
P3.4 Water at 20°C flows through a long elliptical duct 30 cm
wide and 22 cm high. What average velocity, in m/s,
would cause the weight flow to be 500 lbf/s?
P3.5 Water at 20°C flows through a 5-in-diameter smooth
pipe at a high Reynolds number, for which the velocity
profile is approximated by u ≈ Uo(y/R)1/8, where Uo is
the centerline velocity, R is the pipe radius, and y is the
distance measured from the wall toward the centerline. If
the centerline velocity is 25 ft/s, estimate the volume
flow rate in gallons per minute.
The Reynolds transport theorem
P3.6
D3 = 6 cm
d
(mV)
dt
Water fills a cylindrical tank to depth h. The tank has
­diameter D. The water flows out at average velocity Vo
D4 = 9 cm
P3.8
D1 = 4 cm
A laboratory test tank contains seawater of salinity S and
density ρ. Water enters the tank at conditions (S1, ρ1, A1,
V1) and is assumed to mix immediately in the tank. Tank
water leaves through an outlet A2 at velocity V2. If salt is
a “conservative” property (neither created nor destroyed),
use the Reynolds transport theorem to find an expression
for the rate of change of salt mass Msalt within the tank.
P3.9
Conservation of mass
P3.10 Water flowing through an 8-cm-diameter pipe enters a
porous section, as in Fig. P3.10, which allows a uniform
radial velocity vw through the wall surfaces for a distance
of 1.2 m. If the entrance average velocity V1 is 12 m/s,
find the exit velocity V2 if (a) vw = 15 cm/s out of the
pipe walls or (b) vw = 10 cm/s into the pipe. (c) What
value of vw will make V2 = 9 m/s?
vw
V1
V2
1.2 m
D = 8 cm
P3.10
P3.11 Water flows from a faucet into a sink at 3 U.S. gallons
per minute. The stopper is closed, and the sink has two
198
Chapter 3 Integral Relations for a Control Volume
r­ectangular overflow drains, each 3/8 in by 1¼ in. If the
sink water level remains constant, estimate the average
overflow ­velocity, in ft/s.
P3.12 The pipe flow in Fig. P3.12 fills a cylindrical surge tank
as shown. At time t = 0, the water depth in the tank is
30 cm. Estimate the time required to fill the remainder of
the tank.
3
Q3 = 0.01 m 3/s
1
D = 75 cm
2
h
D1 = 5 cm
1m
Water
V1 = 2.5 m/s
d = 12 cm
V2 = 1.9 m/s
d
P3.14
r=R
P3.12
P3.13 The cylindrical container in Fig. P3.13 is 20 cm in diameter and has a conical contraction at the bottom with an
exit hole 3 cm in diameter. The tank contains fresh water
at standard sea-level conditions. If the water surface is
falling at the nearly steady rate dh/dt ≈ −0.072 m/s, estimate the average velocity V out of the bottom exit.
D2 = 7 cm
r
u(r)
U0
x=0
x=L
P3.15
P3.16 An incompressible fluid flows past an impermeable flat
plate, as in Fig. P3.16, with a uniform inlet profile u = U0
and a cubic polynomial exit profile
h(t)
D
u ≈ U0 (
y
3η − η3
where η =
)
2
δ
Compute the volume flow Q across the top surface of the
control volume.
U0
P3.13
y=δ
Q?
U0
V?
P3.14 The open tank in Fig. P3.14 contains water at 20°C and
is being filled through section 1. Assume incompressible
flow. First derive an analytic expression for the waterlevel change dh/dt in terms of arbitrary volume flows
(Q1, Q2, Q3) and tank diameter d. Then, if the water level
h is constant, determine the exit velocity V2 for the given
data V1 = 3 m/s and Q3 = 0.01 m3/s.
P3.15 Water, assumed incompressible, flows steadily through
the round pipe in Fig. P3.15. The entrance velocity is
constant, u = U0, and the exit velocity approximates turbulent flow, u = umax(1 − r/R)1/7. Determine the ratio U0/
umax for this flow.
y=0
Solid plate, width b into paper
CV
Cubic
P3.16
P3.17 Incompressible steady flow in the inlet between parallel
plates in Fig. P3.17 is uniform, u = U0 = 8 cm/s, while
downstream the flow develops into the parabolic laminar
profile u = az(z0 − z), where a is a constant. If z0 = 4 cm
and the fluid is SAE 30 oil at 20°C, what is the value of
umax in cm/s?
Problems 199
D = 10 cm
h = 2 mm
z = z0
u max
U0
z=0
2
P3.17
P3.18 Gasoline enters section 1 in Fig. P3.18 at 0.5 m3/s. It
leaves section 2 at an average velocity of 12 m/s. What is
the average velocity at section 3? Is it in or out?
2
1
P3.20
D1 = 3 mm
(2)
D2 = 18 cm
(1)
Air
D3 = 13 cm
1
P3.18
P3.19 Water from a storm drain flows over an outfall onto a
­porous bed that absorbs the water at a uniform vertical
­velocity of 8 mm/s, as shown in Fig. P3.19. The system
is 5 m deep into the paper. Find the length L of the bed
that will completely absorb the storm water.
Initial depth = 20 cm
2 m/s
L?
D1 = 1 cm
2
D2 = 2.5 cm
P3.22
P3.23 The hypodermic needle in Fig. P3.23 contains a liquid
­serum (SG = 1.05). If the serum is to be injected steadily
at 6 cm3/s, how fast in in/s should the plunger be advanced (a) if leakage in the plunger clearance is neglected
and (b) if leakage is 10 percent of the needle flow?
D1 = 0.75 in
D 2 = 0.030 in
V2
P3.19
P3.20 Oil (SG = 0.89) enters at section 1 in Fig. P3.20 at a
P3.23
weight flow of 250 N/h to lubricate a thrust bearing. The
steady oil flow exits radially through the narrow clearance *P3.24 Water enters the bottom of the cone in Fig. P3.24 at a
between thrust plates. Compute (a) the outlet volume flow
u­ niformly increasing average velocity V = Kt. If d is very
in mL/s and (b) the average outlet velocity in cm/s.
small, derive an analytic formula for the water surface rise h(t)
P3.21 For the two-port tank of Fig. E3.5, assume D1 = 4 cm,
for the condition h = 0 at t = 0. Assume incompressible flow.
V1 = 18 m/s, D2 = 7 cm, and V2 = 8 m/s. If the tank surface is rising at 17 mm/s, estimate the tank diameter.
P3.22 The converging–diverging nozzle shown in Fig. P3.22
Cone
­expands and accelerates dry air to supersonic speeds at
θ
θ
the exit, where p2 = 8 kPa and T2 = 240 K. At the throat,
Diameter d
h(t)
p1 = 284 kPa, T1 = 665 K, and V1 = 517 m/s. For steady
­compressible flow of an ideal gas, estimate (a) the mass
flow in kg/h, (b) the velocity V2, and (c) the Mach numV = Kt
P3.24
ber Ma2.
200
Chapter 3 Integral Relations for a Control Volume
This rate persists as long as p0 is at least twice as large as
P3.25 As will be discussed in Chaps. 7 and 8, the flow of a
the atmospheric pressure. Assuming constant T0 and an
stream U0 past a blunt flat plate creates a broad low-­
ideal gas, (a) derive a formula for the change of density
velocity wake behind the plate. A simple model is given
ρ0(t) within the tank. (b) Analyze the time Δt required for
in Fig. P3.25, with only half of the flow shown due to
symmetry. The velocity profile behind the plate is idealthe density to decrease by 25 percent.
ized as “dead air” (near-zero velocity) behind the plate, P3.28 Air, assumed to be a perfect gas from Table A.4, flows
plus a higher velocity, decaying vertically above the wake
through a long, 2-cm-diameter insulated tube. At section
according to the variation u ≈ U0 + ΔU e−z/L, where L is
1, the pressure is 1.1 MPa and the temperature is 345 K.
At section 2, 67 meters further downstream, the density is
the plate height and z = 0 is the top of the wake. Find ΔU
1.34 kg/m3, the temperature 298 K, and the Mach number
as a function of stream speed U0.
is 0.90. For one-dimensional flow, calculate (a) the mass
U0
flow; (b) p2; (c) V2; and (d ) the change in entropy between
z
1 and 2. (e) How do you explain the entropy change?
Exponential curve
P3.29 In elementary compressible flow theory (Chap. 9), comu
pressed air will exhaust from a small hole in a tank at the
Width b

U0
into paper
mass flow rate m ≈ Cρ, where ρ is the air density in the
U + ∆U
tank and C is a constant. If ρ0 is the initial density in a
tank of volume 𝒱, derive a formula for the density
L
change ρ(t) after the hole is opened. Apply your formula
Dead air (negligible velocity)
2
to the following case: a spherical tank of diameter 50 cm,
with initial pressure 300 kPa and temperature 100°C, and
CL
a hole whose initial exhaust rate is 0.01 kg/s. Find the
P3.25
time required for the tank density to drop by 50 percent.
P3.26 A thin layer of liquid, draining from an inclined plane, as P3.30 For the nozzle of Fig. P3.22, consider the following data
in Fig. P3.26, will have a laminar velocity profile u ≈
for air, k = 1.4. At the throat, p1 = 1000 kPa, V1 = 491
U0(2y/h − y2/h2), where U0 is the surface velocity. If the
m/s, and T1 = 600 K. At the exit, p2 = 28.14 kPa. Assumplane has width b into the paper, determine the volume
ing isentropic steady flow, compute (a) the Mach numrate of flow in the film. Suppose that h = 0.5 in and the
ber Ma1; (b) T2; (c) the mass flow; and (d) V2.
flow rate per foot of channel width is 1.25 gal/min. Esti- P3.31 A bellows may be modeled as a deforming wedgemate U0 in ft/s.
shaped volume as in Fig. P3.31. The check valve on the
left (pleated) end is closed during the stroke. If b is the
g
y
bellows width into the paper, derive an expression for

outlet mass flow m0 as a function of stroke θ(t).
L
h
u (y)
h
θ
θ (t)
P3.26
P3.27 Consider a highly pressurized air tank at conditions (p0,
ρ0, T0) and volume υ0. In Chap. 9 we will learn that, if the
tank is allowed to exhaust to the atmosphere through a
well-designed converging nozzle of exit area A, the outgoing mass flow rate will be
α p0 A

m=
where α ≈ 0.685 for air
√RT0
m0
θ (t)
x
h
Stroke
P3.31
d≪h
Problems 201
P3.32 Water at 20°C flows steadily through the piping junction
in Fig. P3.32, entering section 1 at 20 gal/min. The average velocity at section 2 is 2.5 m/s. A portion of the flow
is ­diverted through the showerhead, which contains 100
holes of 1-mm diameter. Assuming uniform shower
flow, estimate the exit velocity from the showerhead jets.
d = 4 cm
(3)
Propellant
d = 1.5 cm
d = 2 cm
P3.32
(2)
Exit section
De = 18 cm
pe = 90 kPa
Ve = 1150 m/s
Te = 750 K
Combustion:
1500 K, 950 kPa
(1)
Propellant
P3.33 In some wind tunnels the test section is perforated to
suck out fluid and provide a thin viscous boundary layer.
The test section wall in Fig. P3.33 contains 1200 holes of
5-mm diameter each per square meter of wall area. The
suction velocity through each hole is Vs = 8 m/s, and the
test-­section entrance velocity is V1 = 35 m/s. Assuming
incompressible steady flow of air at 20°C, compute (a)
V0, (b) V2, and (c) Vf, in m/s.
Df = 2.2 m
Test section
Ds = 0.8 m
Uniform suction
V2
Vf
P3.35 In contrast to the liquid rocket in Fig. P3.34, the solid-­
propellant rocket in Fig. P3.35 is self-contained and has
no entrance ducts. Using a control volume analysis for
the conditions shown in Fig. P3.35, compute the rate of
mass loss of the propellant, assuming that the exit gas
has a ­molecular weight of 28.
D0 = 2.5 m
V1
V0
P3.35
P3.36 The jet pump in Fig. P3.36 injects water at U1 = 40 m/s
through a 3-in pipe and entrains a secondary flow of
water U2 = 3 m/s in the annular region around the small
pipe. The two flows b­ ecome fully mixed downstream,
where U3 is approximately constant. For steady incompressible flow, compute U3 in m/s.
D1 = 3 in
Mixing
region
Inlet
U1
L=4m
U3
U2
P3.33
D2 = 10 in
P3.34 A rocket motor is operating steadily, as shown in
Fig. P3.34. The products of combustion flowing out the
exhaust nozzle approximate a perfect gas with a molecular
weight of 28. For the given conditions calculate V2 in ft/s.
1
Liquid oxygen:
0.5 slug/s
2
4000° R
400 lbf/in 2
15 lbf/in 2
1100° F
D 2 = 5.5 in
3
P3.34
Fully
mixed
Liquid fuel:
0.1 slug/s
P3.36
P3.37 If the rectangular tank full of water in Fig. P3.37 has its
right-hand wall lowered by an amount δ, as shown, water will flow out as it would over a weir or dam. In
Prob. P1.14 we deduced that the outflow Q would be
given by
Q = C b g1/2 δ 3/2
where b is the tank width into the paper, g is the acceleration of gravity, and C is a dimensionless constant. Assume that the water surface is horizontal, not
slightly curved as in the figure. Let the initial excess
water level be δo. Derive a formula for the time required to reduce the excess water level to (a) δo/10 and
(b) zero.
202
Chapter 3 Integral Relations for a Control Volume
δ
Plate
Q ∝ δ 3/2
Dj = 10 cm
h
Vj = 8 m/s
F
L
P3.37
P3.40
P3.38 An incompressible fluid in Fig. P3.38 is being squeezed
outward between two large circular disks by the uniform
downward motion V0 of the upper disk. Assuming one-­ P3.41 In Fig. P3.41 the vane turns the water jet completely
dimensional radial outflow, use the control volume
around. Find an expression for the maximum jet velocity
shown to derive an expression for V(r).
V0 if the maximum possible support force is F0.
V0
CV
h(t)
ρ 0 , V0 , D0
CV
r
V
F0
V(r)?
P3.41
Fixed circular disk
P3.42 A liquid of density ρ flows through the sudden contraction in Fig. P3.42 and exits to the atmosphere. Assume
uniform conditions (p1, V1, D1) at section 1 and (p2, V2,
D2) at ­section 2. Find an expression for the force F exerted by the fluid on the contraction.
P3.38
The linear momentum equation
P3.39 A wedge splits a sheet of 20°C water, as shown in Fig.
P3.39. Both wedge and sheet are very long into the paper. If the force required to hold the wedge stationary is
F = 124 N per meter of depth into the paper, what is the
angle θ of the wedge?
Atmosphere
p1
pa
6 m/s
2
6 m/s
θ
F
P3.42
1
4 cm
6 m/s
P3.39
P3.40 The water jet in Fig. P3.40 strikes normal to a fixed
plate. Neglect gravity and friction, and compute the
force F in newtons required to hold the plate fixed.
P3.43 Water at 20°C flows through a 5-cm-diameter pipe
that has a 180° vertical bend, as in Fig. P3.43. The
total length of pipe between flanges 1 and 2 is 75 cm.
When the weight flow rate is 230 N/s, p1 = 165 kPa
and p2 = 134 kPa. ­Neglecting pipe weight, determine
the total force that the flanges must withstand for this
flow.
Problems 203
2
1
P3.43
*P3.44 When a uniform stream flows past an immersed thick
cylinder, a broad low-velocity wake is created downstream, idealized as a V shape in Fig. P3.44. Pressures p1
and p2 are approximately equal. If the flow is two-­
dimensional and incompressible, with width b into the
paper, derive a formula for the drag force F on the cylinder. Rewrite your ­result in the form of a dimensionless
drag coefficient based on body length CD = F/(ρU2bL).
being a fraction. The reason is that for frictionless flow the
fluid can exert no tangential force Ft on the plate. The
condition Ft = 0 enables us to solve for α. Perform this
analysis, and find α as a function of the plate angle θ. Why
doesn’t the answer depend on the properties of the jet?
P3.47 A liquid jet of velocity Vj and diameter Dj strikes a fixed
hollow cone, as in Fig. P3.47, and deflects back as a
conical sheet at the same velocity. Find the cone angle θ
for which the restraining force F = 32 ρAjV2j .
P3.48 The small boat in Fig. P3.48 is driven at a steady speed V0
by a jet of compressed air issuing from a 3-cm-diameter
hole at Ve = 343 m/s. Jet exit conditions are pe = 1 atm and
Te = 30°C. Air drag is negligible, and the hull drag is kV20,
where k ≈ 19 N · s2/m2. Estimate the boat speed V0 in m/s.
αQ, V
2
ρ , Q, A, V
θ
U
U
1
Fn
U
2
2L
1
U
Ft = 0
L
P3.46
(1-α) Q, V
L
3
Conical sheet
2
Jet
P3.44
θ
F
P3.45 Water enters and leaves the 6-cm-diameter pipe bend in
Fig. P3.45 at an average velocity of 8.5 m/s. The horizontal force to support the bend against momentum
change is 300 N. Find (a) the angle ϕ; and (b) the vertical
force on the bend.
P3.47
ϕ
De = 3 cm
Ve
Compressed
air
V0
P3.45
P3.46 When a jet strikes an inclined fixed plate, as in Fig. P3.46,
it breaks into two jets at 2 and 3 of equal velocity V = Vjet
but unequal flows αQ at 2 and (1 − α)Q at ­section 3, α
Hull drag kV02
P3.48
204
Chapter 3 Integral Relations for a Control Volume
P3.49 The horizontal nozzle in Fig. P3.49 has D1 = 12 in and
D2 = 6 in, with inlet pressure p1 = 38 lbf/in2 absolute and
V2 = 56 ft/s. For water at 20°C, compute the horizontal
force provided by the flange bolts to hold the nozzle fixed.
pa = 15 lbf/in2 abs
Open
jet
Water
P3.52 A large commercial power washer delivers 21 gal/min of
water through a nozzle of exit diameter one-third of an
inch. Estimate the force of the water jet on a wall normal
to the jet.
P3.53 Consider incompressible flow in the entrance of a circular tube, as in Fig. P3.53. The inlet flow is uniform, u1 =
U0. The flow at section 2 is developed pipe flow. Find
the wall drag force F as a function of (p1, p2, ρ, U0, R) if
the flow at section 2 is
(a) Laminar: u2 = umax 1 −
(
2
(b) Turbulent: u2 ≈ umax (1 −
1
P3.49
r2
R2)
P3.50 The jet engine on a test stand in Fig. P3.50 admits air at
20°C and 1 atm at section 1, where A1 = 0.5 m2 and V1 =
250 m/s. The fuel-to-air ratio is 1:30. The air leaves section
2 at atmospheric pressure and higher temperature, where
V2 = 900 m/s and A2 = 0.4 m2. Compute the horizontal
test stand reaction Rx needed to hold this engine fixed.
r 1/7
R)
2
r=R
1
U0
r
x
m fuel
Friction drag on fluid
P3.53
1
P3.50
Combustion
chamber
2
Rx
P3.51 A liquid jet of velocity Vj and area Aj strikes a single 180°
bucket on a turbine wheel rotating at angular velocity Ω,
as in Fig. P3.51. Derive an expression for the power P
delivered to this wheel at this instant as a function of the
system parameters. At what angular velocity is the maximum power delivered? How would your analysis differ if
there were many, many buckets on the wheel, so that the
jet was continually striking at least one bucket?
P3.54 For the pipe-flow-reducing section of Fig. P3.54, D1 =
8 cm, D2 = 5 cm, and p2 = 1 atm. All fluids are at
20°C. If V1 = 5 m/s and the manometer reading is h =
58 cm, estimate the total force resisted by the flange
bolts.
1
2
p2 ≈ pa = 101 kPa
Water
h
Mercury
Bucket
Wheel, radius R
Jet
Ω
P3.51
P3.54
P3.55 In Fig. P3.55 the jet strikes a vane that moves to the right
at constant velocity Vc on a frictionless cart. Compute (a)
the force Fx required to restrain the cart and (b) the power
P delivered to the cart. Also find the cart velocity for
which (c) the force Fx is a maximum and (d) the power P
is a maximum.
Problems 205
θ
8 m/s
60°
D=4m
ρ, Vj , Aj
D0 = 4 cm
Vc = constant
Cable
Fy
Fx
P3.55
P3.58
P3.56 Water at 20°C flows steadily through the box in Fig.
P3.56, entering station (1) at 2 m/s. Calculate the (a)
horizontal and (b) vertical forces required to hold the
box stationary against the flow momentum.
P3.57 Water flows through the duct in Fig. P3.57, which is 50
cm wide and 1 m deep into the paper. Gate BC completely closes the duct when β = 90°. Assuming onedimensional flow, for what angle β will the force of the
exit jet on the plate be 3 kN?
P3.59 When a pipe flow suddenly expands from A1 to A2, as in
Fig. P3.59, low-speed, low-friction eddies appear in the corners and the flow gradually expands to A2 downstream. Using the suggested control volume for incompressible steady
flow and assuming that p ≈ p1 on the corner annular ring as
shown, show that the downstream pressure is given by
p2 = p1 + ρV21
A1
A1
1− )
A2 (
A2
Neglect wall friction.
Control
volume
Pressure ≈ p1
D1 = 5 cm
65˚
D2 = 3 cm
p2 , V2 , A 2
y
P3.56
P3.59
P3.60 Water at 20°C flows through the elbow in Fig. P3.60 and
exits to the atmosphere. The pipe diameter is D1 = 10 cm,
while D2 = 3 cm. At a weight flow rate of 150 N/s, the pressure p1 = 2.3 atm (gage). Neglecting the weight of water and
elbow, estimate the force on the flange bolts at section 1.
Hinge B
1.2 m/s
β
50 cm
p1 , V1 , A1
x
C
F = 3 kN
1
P3.57
P3.58 The water tank in Fig. P3.58 stands on a frictionless cart
and feeds a jet of diameter 4 cm and velocity 8 m/s,
which is deflected 60° by a vane. Compute the tension in
the ­supporting cable.
40°
P3.60
2
206
Chapter 3 Integral Relations for a Control Volume
P3.61 A 20°C water jet strikes a vane mounted on a tank with
frictionless wheels, as in Fig. P3.61. The jet turns
and falls into the tank without spilling out. If θ = 30°,
evaluate the horizontal force F required to hold the tank
stationary.
passes through the hole, and part is deflected. Determine
the horizontal force required to hold the plate.
Plate
Vj = 50 ft/s
D1 = 6 cm
θ
Dj = 2 in
D2 = 4 cm
25 m/s
Water
25 m/s
F
P3.61
P3.62 Water at 20°C exits to the standard sea-level atmosphere
through the split nozzle in Fig. P3.62. Duct areas are A1
= 0.02 m2 and A2 = A3 = 0.008 m2. If p1 = 135 kPa (absolute) and the flow rate is Q2 = Q3 = 275 m3/h, compute
the force on the flange bolts at section 1.
2
P3.64
P3.65 The box in Fig. P3.65 has three 0.5-in holes on the right
side. The volume flows of 20°C water shown are steady,
but the details of the interior are not known. Compute
the force, if any, that this water flow causes on the box.
30°
0.1 ft3 /s
0.2 ft3 /s
30°
0.1 ft3 /s
1
3
P3.62
P3.65
P3.63 Water flows steadily through the box in Fig. P3.63.
­Average velocity at all ports is 7 m/s. The vertical momentum force on the box is 36 N. What is the inlet mass
flow?
(2)
40°
P3.66 The tank in Fig. P3.66 weighs 500 N empty and contains
600 L of water at 20°C. Pipes 1 and 2 have equal diameters of 6 cm and equal steady volume flows of 300 m3/h.
What should the scale reading W be in N?
(3)
1
W?
40°
2
(1)
Water
P3.63
P3.64 The 6-cm-diameter 20°C water jet in Fig. P3.64 strikes a
plate containing a hole of 4-cm diameter. Part of the jet
Scale
P3.66
Problems 207
P3.67 For the boundary layer of Fig. 3.9, for air, ρ = 1.2 kg/m3,
let h = 7 cm, Uo = 12 m/s, b = 2 m, and L = 1 m. Let the
velocity at the exit, x = L, approximate a turbulent flow:
u/Uo ≈ (y/δ) 1/7 . Calculate (a) δ and (b) the friction
drag D.
P3.68 The rocket in Fig. P3.68 has a supersonic exhaust, and
the exit pressure pe is not necessarily equal to pa. Show
that the force F required to hold this rocket on the test
stand is F = ρe AeV2e + Ae ( pe − pa ). Is this force F what
we term the thrust of the rocket?
Fuel
.
mf
p a ≠ pe
F
pe , Ae ,Ve
.
m0
e
Oxidizer
P3.68
P3.69 A uniform rectangular plate, 40 cm long and 30 cm deep
into the paper, hangs in air from a hinge at its top (the
­30-cm side). It is struck in its center by a horizontal
­3-cm-diameter jet of water moving at 8 m/s. If the gate
has a mass of 16 kg, estimate the angle at which the plate
will hang from the vertical.
P3.70 The dredger in Fig. P3.70 is loading sand (SG = 2.6)
onto a barge. The sand leaves the dredger pipe at
4 ft/s with a weight flow of 850 lbf/s. Estimate the
tension on the mooring line caused by this loading
process.
30°
45°
45°
P3.71
*P3.72 When immersed in a uniform stream, a thick elliptical
­cylinder creates a broad downstream wake, as idealized
in Fig. P3.72. The pressure at the upstream and downstream sections are approximately equal, and the fluid is
water at 20°C. If U0 = 4 m/s and L = 80 cm, estimate the
drag force on the cylinder per unit width into the paper.
Also compute the dimensionless drag coefficient CD =
2F/(ρU20bL).
P3.73 A pump in a tank of water at 20°C directs a jet at 45 ft/s
and 200 gal/min against a vane, as shown in Fig. P3.73.
Compute the force F to hold the cart stationary if the jet
follows (a) path A or (b) path B. The tank holds 550 gal
of water at this instant.
U0
U0
L
U0
2
Width b into paper
P3.72
B
A
120°
60°
P3.70
P3.71 Suppose that a deflector is deployed at the exit of the jet
engine of Prob. P3.50, as shown in Fig. P3.71. What
will the reaction Rx on the test stand be now? Is this reaction sufficient to serve as a braking force during airplane
­landing?
F
Water
P3.73
L
L
208
Chapter 3 Integral Relations for a Control Volume
P3.74 Water at 20°C flows down through a vertical, 6-cm-­
diameter tube at 300 gal/min, as in Fig. P3.74. The flow
then turns horizontally and exits through a 90° radial
duct segment 1 cm thick, as shown. If the radial outflow
is uniform and steady, estimate the forces (Fx, Fy, Fz)
required to support this system against fluid momentum
changes.
Horizontal
plane
z
pa = 100 kPa
x
90°
x
1 cm
1
y
6 cm
Vertical
plane
kPa, D1 = 25 cm, V1 = 2.2 m/s, p2 = 120 kPa, and D2
= 8 cm. N
­ eglecting bend and water weight, estimate
the total force that must be resisted by the flange
bolts.
R = 15 cm
Radial outflow
P3.74
*P3.75 A jet of liquid of density ρ and area A strikes a block and
splits into two jets, as in Fig. P3.75. Assume the same
­velocity V for all three jets. The upper jet exits at an angle θ and area αA. The lower jet is turned 90° downward.
­Neglecting fluid weight, (a) derive a formula for the
forces (Fx, Fy) required to support the block against fluid
­momentum changes. (b) Show that Fy = 0 only if α ≥ 0.5.
(c) Find the values of α and θ for which both Fx and Fy are
zero.
αA
ρ
θ
2
P3.77
P3.78 A fluid jet of diameter D1 enters a cascade of moving
blades at absolute velocity V1 and angle β1, and it leaves
at absolute velocity V2 and angle β2, as in Fig. P3.78. The
blades move at velocity u. Derive a formula for the
power P delivered to the blades as a function of these
parameters.
P3.79 The Saturn V rocket in the chapter opener photo was
powered by five F-1 engines, each of which burned 3945
lbm/s of liquid oxygen and 1738 lbm of kerosene per second. The exit velocity of burned gases was approximately
8500 ft/s. In the spirit of Prob. P3.34, neglecting external
pressure forces, estimate the total thrust of the rocket, in
lbf.
V, A
Fx
α2
α1
u
β2
V2
(1 – α)A
P3.75
V1
Fy
P3.76 A two-dimensional sheet of water, 10 cm thick and moving at 7 m/s, strikes a fixed wall inclined at 20° with respect to the jet direction. Assuming frictionless flow,
find (a) the normal force on the wall per meter of depth,
and find the widths of the sheet deflected (b) upstream
and (c) downstream along the wall.
P3.77 Water at 20°C flows steadily through a reducing pipe
bend, as in Fig. P3.77. Known conditions are p1 = 350
P3.78
Air jet
D1
β1
Blades
P3.80 A river of width b and depth h1 passes over a submerged
obstacle, or “drowned weir,” in Fig. P3.80, emerging at
a new flow condition (V2, h2). Neglect atmospheric
pressure, and assume that the water pressure is hydro-
Problems 209
static at both sections 1 and 2. Derive an expression for
the force exerted by the river on the obstacle in terms of
V1, h1, h2, b, ρ, and g. Neglect water friction on the river
bottom.
Width b into paper
V1, h1
V2, h2
P3.83 Gasoline at 20°C is flowing at V1 = 12 m/s in a 5-cm-­
diameter pipe when it encounters a 1-m length of uniform radial wall suction. At the end of this suction
region, the average fluid velocity has dropped to V2 = 10
m/s. If p1 = 120 kPa, estimate p2 if the wall friction
losses are ­neglected.
P3.84 Air at 20°C and 1 atm flows in a 25-cm-diameter duct at
15 m/s, as in Fig. P3.84. The exit is choked by a 90°
cone, as shown. Estimate the force of the airflow on the
cone.
P3.80
1 cm
P3.81 Torricelli’s idealization of efflux from a hole in the side
of a tank is V = √2 gh, as shown in Fig. P3.81. The cylindrical tank weighs 150 N when empty and contains
water at 20°C. The tank bottom is on very smooth ice
(static friction coefficient ζ ≈ 0.01). The hole diameter is
9 cm. For what water depth h will the tank just begin to
move to the right?
25 cm
90°
40 cm
P3.84
Water
h
V
P3.81
1m
P3.85 The thin-plate orifice in Fig. P3.85 causes a large pressure drop. For 20°C water flow at 500 gal/min, with pipe
D = 10 cm and orifice d = 6 cm, p1 − p2 ≈ 145 kPa. If
the wall friction is negligible, estimate the force of the
water on the orifice plate.
30 cm
Static
friction
*P3.82 The model car in Fig. P3.82 weighs 17 N and is to be
­accelerated from rest by a 1-cm-diameter water jet moving at 75 m/s. Neglecting air drag and wheel friction,
estimate the velocity of the car after it has moved forward 1 m.
x
Vj
V
P3.82
P3.85
1
2
P3.86 For the water jet pump of Prob. P3.36, add the following
data: p1 = p2 = 25 lbf/in2, and the distance between sections 1 and 3 is 80 in. If the average wall shear stress
­between sections 1 and 3 is 7 lbf/ft2, estimate the pressure p3. Why is it higher than p1?
P3.87 A vane turns a water jet through an angle α, as shown
in Fig. P3.87. Neglect friction on the vane walls. (a)
What is the angle α for the support force to be in pure
210
Chapter 3 Integral Relations for a Control Volume
pa
compression? (b) Calculate this compression force if
the water velocity is 22 ft/s and the jet cross section is
4 in2.
V
h
F
P3.87
V(t)
V
α
Stopper
P3.90
25°
P3.88 The boat in Fig. P3.88 is jet-propelled by a pump that
­develops a volume flow rate Q and ejects water out the
stern at velocity Vj. If the boat drag force is F = kV2,
where k is a constant, develop a formula for the steady
forward speed V of the boat.
*P3.92 A more involved version of Prob. P3.90 is the elbowshaped tube in Fig. P3.92, with constant cross-sectional
area A and diameter D ≪ h, L. Assume incompressible
flow, neglect friction, and derive a differential equation
for dV/dt when the stopper is opened. Hint: Combine
two ­control volumes, one for each leg of the tube.
pa
V
Pump
Q
Vj
h
V1
P3.88
P3.89 Consider Fig. P3.36 as a general problem for analysis
of a mixing ejector pump. If all conditions (p, ρ, V) are
known at sections 1 and 2 and if the wall friction is
negligible, derive formulas for estimating (a) V3 and
(b) p3.
P3.90 As shown in Fig. P3.90, a liquid column of height h is
­confined in a vertical tube of cross-sectional area A
by a stopper. At t = 0 the stopper is suddenly removed, exposing the bottom of the liquid to atmospheric pressure. Using a control volume analysis of
mass and vertical momentum, derive the differential
equation for the downward motion V(t) of the liquid.
Assume one-dimensional, incompressible, frictionless flow.
P3.91 Extend Prob. P3.90 to include a linear (laminar) average
wall shear stress resistance of the form τ ≈ cV, where c is
a constant. Find the differential equation for dV/dt and
then solve for V(t), assuming for simplicity that the wall
area remains constant.
L
V2
P3.92
P3.93 According to Torricelli’s theorem, the velocity of a fluid
draining from a hole in a tank is V ≈ (2gh)1/2, where h is
the depth of water above the hole, as in Fig. P3.93. Let
the hole have area Ao and the cylindrical tank have crosssection area Ab ≫ Ao. Derive a formula for the time to
drain the tank completely from an initial depth ho.
Water
P3.93
V
h
Problems 211
Neglect the short horizontal leg, and combine control
­volume analyses for the left and right legs to derive a
single differential equation for V(t) of the liquid
­column.
*P3.97 Extend Prob. P3.96 to include a linear (laminar) average
wall shear stress resistance of the form τ ≈ 8µV/D, where
3 in
µ is the fluid viscosity. Find the differential equation for
dV/dt and then solve for V(t), assuming an initial disV
placement z = z0, V = 0 at t = 0. The result should be a
36 in
damped oscillation tending toward z = 0.
*P3.98 As an extension of Example 3.11, let the plate and its
cart (see Fig. 3.10a) be unrestrained horizontally, with
20 in
frictionless wheels. Derive (a) the equation of motion for
cart ­velocity Vc(t) and (b) a formula for the time required
for the cart to accelerate from rest to 90 percent of the jet
velocity (assuming the jet continues to strike the plate
8 in
P3.94
horizontally). (c) Compute numerical values for part (b)
P3.95 A tall water tank discharges through a well-rounded
using the conditions of Example 3.11 and a cart mass of
­orifice, as in Fig. P3.95. Use the Torricelli formula of
2 kg.
Prob. P3.81 to estimate the exit velocity. (a) If, at this P3.99 Let the rocket of Fig. E3.12 start at z = 0, with constant
instant, the force F required to hold the plate is 40 N,
exit velocity and exit mass flow, and rise vertically with
what is the depth h? (b) If the tank surface is dropping at
zero drag. (a) Show that, as long as fuel burning continthe rate of 2.5 cm/s, what is the tank diameter D?
ues, the vertical height S(t) reached is given by

Ve Mo
mt
S=
 [ζlnζ − ζ + 1], where ζ = 1 −
m
Mo
P3.94 A water jet 3 in in diameter strikes a concrete (SG = 2.3)
slab which rests freely on a level floor. If the slab is 1 ft
wide into the paper, calculate the jet velocity which will
just begin to tip the slab over.
h
F
d = 4 cm
D
P3.95
P3.96 Extend Prob. P3.90 to the case of the liquid motion in a
frictionless U-tube whose liquid column is displaced a
distance Z upward and then released, as in Fig. P3.96.
z
z
Equilibrium position
Liquid – column length
L = h1 + h2 + h3
h1
(b) Apply this to the case Ve = 1500 m/s and Mo =
1000 kg to find the height reached after a burn of 30
seconds, when the final rocket mass is 400 kg.
P3.100 Suppose that the solid-propellant rocket of Prob. P3.35 is
built into a missile of diameter 70 cm and length 4 m.
The system weighs 1800 N, which includes 700 N of
propellant. Neglect air drag. If the missile is fired vertically from rest at sea level, estimate (a) its velocity and
height at fuel burnout and (b) the maximum height it will
attain.
P3.101 Water at 20°C flows steadily through the tank in Fig.
P3.101. Known conditions are D1 = 8 cm, V1 = 6 m/s,
and D2 = 4 cm. A rightward force F = 70 N is required
to keep the tank fixed. (a) What is the velocity leaving
section 2? (b) If the tank cross section is 1.2 m2, how fast
is the water surface h(t) rising or falling?
h3
V
2
P3.96
h2 ≈ 0
P3.101
F
h (t)
1
212
Chapter 3 Integral Relations for a Control Volume
P3.102 As can often be seen in a kitchen sink when the faucet is
running, a high-speed channel flow (V1, h1) may “jump” to
a low-speed, low-energy condition (V2, h2) as in Fig. P3.102.
The pressure at sections 1 and 2 is approximately hydrostatic, and wall friction is negligible. Use the continuity and
momentum relations to find h2 and V2 in terms of (h1, V1).
which is turned completely around by the parachute, as
in Fig. P3.106. (a) Find the force F required to support
the chute. (b) Express this force as a dimensionless drag
­coefficient, CD = F/[(½)ρV2(π/4)D2] and compare with
Table 7.3.
Hydraulic
jump
ρ, V
V2 < V1
h1
h2 > h1
V1
D/2
P3.106
P3.102
*P3.103 Suppose that the solid-propellant rocket of Prob. P3.35 is
mounted on a 1000-kg car to propel it up a long slope of 15°.
The rocket motor weighs 900 N, which includes 500 N of
propellant. If the car starts from rest when the rocket is fired,
and if air drag and wheel friction are neglected, ­estimate the
maximum distance that the car will travel up the hill.
P3.104 A rocket is attached to a rigid horizontal rod hinged at
the origin as in Fig. P3.104. Its initial mass is M0, and its

exit properties are m and Ve relative to the rocket. Set up
the differential equation for rocket motion, and solve for
the angular velocity ω(t) of the rod. Neglect gravity, air
drag, and the rod mass.
x
P3.107 The cart in Fig. P3.107 moves at constant velocity V0 =
12 m/s and takes on water with a scoop 80 cm wide that dips
h = 2.5 cm into a pond. Neglect air drag and wheel friction.
Estimate the force required to keep the cart moving.
V0
Water
h
P3.107
*P3.108 A rocket sled of mass M is to be decelerated by a scoop, as
in Fig. P3.108, which has width b into the paper and dips
into the water a depth h, creating an upward jet at 60°. The
rocket thrust is T to the left. Let the initial velocity be V0,
and neglect air drag and wheel friction. Find an expression
for V(t) of the sled for (a) T = 0 and (b) finite T ≠ 0.
R
y
∙
ω, ω
P3.104
D
.
m, Ve , pe = pa
P3.105 Extend Prob. P3.104 to the case where the rocket has a
linear air drag force F = cV, where c is a constant. Assuming no burnout, solve for ω(t) and find the terminal
angular velocity—that is, the final motion when the angular acceleration is zero. Apply to the case M0 = 6 kg,

R = 3 m, m = 0.05 kg/s, Ve = 1100 m/s, and c = 0.075
N · s/m to find the angular velocity after 12 s of burning.
P3.106 Actual airflow past a parachute creates a variable distribution of velocities and directions. Let us model this as a
­circular air jet, of diameter half the parachute diameter,
60°
M
Water
P3.108
h
V
Problems 213
P3.109 For the boundary-layer flow in Fig. 3.9, let the exit velocity profile, at x = L, simulate turbulent flow,
u ≈ U0 (y/δ) 1/7 . (a) Find a relation between h and δ.
(b) Find an expression for the drag force F on the plate
between 0 and L.
D1 = 10 cm
D2 = 3 cm
F
Water at 20°C
The Bernoulli Equation
P3.110 Repeat Prob. P3.49 by assuming that p1 is unknown
and using Bernoulli’s equation with no losses. Compute the new bolt force for this assumption. What is
the head loss between 1 and 2 for the data of Prob.
P3.49?
P3.111 As a simpler approach to Prob. P3.96, apply the unsteady
Bernoulli equation between 1 and 2 to derive a differential equation for the motion z(t). Neglect friction and
compressibility.
P3.112 A jet of alcohol strikes the vertical plate in Fig. P3.112.
A force F ≈ 425 N is required to hold the plate stationary. ­Assuming there are no losses in the nozzle, estimate
(a) the mass flow rate of alcohol and (b) the absolute
pressure at section 1.
Alcohol , SG = 0.79
Air
h?
Hg
P3.114
P3.115 A free liquid jet, as in Fig. P3.115, has constant ambient
pressure and small losses; hence from Bernoulli’s equation z + V2/(2g) is constant along the jet. For the fire
nozzle in the figure, what are (a) the minimum and (b)
the maximum values of θ for which the water jet will
clear the corner of the building? For which case will the
jet velocity be higher when it strikes the roof of the
building?
X
pa = 101 kPa
–V2
V1
V1 = 100 ft/s
F
D2 = 2 cm
D1 = 5 cm
P3.112
P3.113 An airplane is flying at 300 mi/h at 4000 m standard
­altitude. As is typical, the air velocity relative to the upper surface of the wing, near its maximum thickness, is
26 percent higher than the plane’s velocity. Using Bernoulli’s equation, calculate the absolute pressure at this
point on the wing. Neglect elevation changes and compressibility.
P3.114 Water flows through a circular nozzle, exits into the air as a
jet, and strikes a plate, as shown in Fig. P3.114. The force
required to hold the plate steady is 70 N. Assuming steady,
frictionless, one-dimensional flow, estimate (a) the velocities at sections (1) and (2) and (b) the mercury manometer
reading h.
50 ft
θ
40 ft
P3.115
P3.116 For the container of Fig. P3.116 use Bernoulli’s equation to derive a formula for the distance X where the
free jet leaving horizontally will strike the floor, as a
function of h and H. For what ratio h/H will X be maximum? Sketch the three trajectories for h/H = 0.25, 0.5,
and 0.75.
Free
jet
H
h
P3.116
X
214
Chapter 3 Integral Relations for a Control Volume
P3.117 Water at 20°C, in the pressurized tank of Fig. P3.117,
flows out and creates a vertical jet as shown. Assuming
steady frictionless flow, determine the height H to which
the jet rises.
3 in
1 in
Air
75 kPa (gage)
P3.120
H?
Water
85 cm
P3.117
P3.118 Bernoulli’s 1738 treatise Hydrodynamica contains many
excellent sketches of flow patterns related to his frictionless relation. One, however, redrawn here as Fig. P3.118,
seems physically misleading. Can you explain what
might be wrong with the figure?
P3.121 In Fig. P3.121 the flowing fluid is CO2 at 20°C. Neglect
losses. If p1 = 170 kPa and the manometer fluid is Meriam
red oil (SG = 0.827), estimate (a) p2 and (b) the gas flow
rate in m3/h.
P3.122 The cylindrical water tank in Fig. P3.122 is being filled
at a volume flow Q1 = 1.0 gal/min, while the water also
drains from a bottom hole of diameter d = 6 mm. At time
t = 0, h = 0. Find and plot the variation h(t) and the eventual maximum water depth hmax. Assume that Bernoulli’s
steady-flow equation is valid.
D2 = 6 cm
D1 = 10 cm
8 cm
P3.121
Jet
Jet
P3.118
P3.119 A long fixed tube with a rounded nose, aligned with an
oncoming flow, can be used to measure velocity. Measurements are made of the pressure at (1) the front nose
and (2) a hole in the side of the tube further along, where
the pressure nearly equals stream pressure.
(a) Make a sketch of this device and show how the velocity is calculated. (b) For a particular sea-level airflow,
the difference between nose pressure and side pressure is
1.5 lbf/in2. What is the air velocity, in mi/h?
P3.120 The manometer fluid in Fig. P3.120 is mercury. Estimate
the volume flow in the tube if the flowing fluid is (a)
gasoline and (b) nitrogen, at 20°C and 1 atm.
Q1
Diameter
D = 20 cm
h
P3.122
V2
P3.123 The air-cushion vehicle in Fig. P3.123 brings in sea-level
standard air through a fan and discharges it at high velocity through an annular skirt of 3-cm clearance. If the
vehicle weighs 50 kN, estimate (a) the required airflow
rate and (b) the fan power in kW.
Problems 215
W = 50 kN
Air:
p = 20 lbf/in2 abs
pa = 14.7 lbf/in2 abs
h = 3 cm
5 ft
1
D = 1 in
V
V
P3.126
D=6m
P3.127 In Fig. P3.127 the open jet of water at 20°C exits a nozzle into sea-level air and strikes a stagnation tube as
shown.
P3.123
P3.124 A necked-down section in a pipe flow, called a venturi,
develops a low throat pressure that can aspirate fluid
upward from a reservoir, as in Fig. P3.124. Using
­
­Bernoulli’s equation with no losses, derive an expression
for the velocity V1 that is just sufficient to bring reservoir
fluid into the throat.
(1)
Open jet
P3.127
D1
V1
12 cm
H
Sea-level air
D2
Water
4 cm
Water
V2, p2 = pa
h
pa
Water
P3.124
P3.125 Suppose you are designing an air hockey table. The
table is 3.0 × 6.0 ft in area, with 161 -in-diameter holes
spaced every inch in a rectangular grid pattern (2592
holes total). The required jet speed from each hole is
estimated to be 50 ft/s. Your job is to select an appropriate blower that will meet the requirements. Estimate
the volumetric flow rate (in ft3/min) and pressure rise
(in lb/in2) required of the blower. Hint: Assume that
the air is stagnant in the large volume of the manifold
under the table surface, and neglect any frictional
losses.
P3.126 The liquid in Fig. P3.126 is kerosene at 20°C. Estimate
the flow rate from the tank for (a) no losses and (b) pipe
losses hf ≈ 4.5V2/(2g).
If the pressure at the centerline at section 1 is 110 kPa, and
losses are neglected, estimate (a) the mass flow in kg/s
and (b) the height H of the fluid in the stagnation tube.
P3.128 A venturi meter, shown in Fig. P3.128, is a carefully
­designed constriction whose pressure difference is a
measure of the flow rate in a pipe. Using Bernoulli’s
equation for steady incompressible flow with no losses,
show that the flow rate Q is related to the manometer
reading h by
Q=
− (D2/D1 ) 4 √
A2
√1
2gh(ρM − ρ)
ρ
where ρM is the density of the manometer fluid.
1
2
h
P3.128
216
Chapter 3 Integral Relations for a Control Volume
P3.129 A water stream flows past a small circular cylinder at
23 ft/s, approaching the cylinder at 3000 lbf/ft2. Measurements at low (laminar flow) Reynolds numbers indicate a maximum surface velocity 60 percent higher than
the stream velocity at point B on the cylinder. Estimate
the pressure at B.
P3.130 In Fig. P3.130 the fluid is gasoline at 20°C at a weight
flow of 120 N/s. Assuming no losses, estimate the gage
pressure at section 1.
approximates turbulent flow in a 2-m-long tube. Calculate the exit velocity in m/s and the volume flow rate in
cm3/s, and compare to Example 3.14.
P3.133 If losses are neglected in Fig. P3.133, for what water
level h will the flow begin to form vapor cavities at the
throat of the nozzle?
pa = 100 kPa
5 cm
D1 = 5 cm
h
p1
12 m
D2 = 8 cm
Open
jet
Open
jet
2
1
Water at 30°C
2
P3.133
8 cm
*P3.134 For the 40°C water flow in Fig. P3.134, estimate the
­volume flow through the pipe, assuming no losses;
then explain what is wrong with this seemingly inP3.131 In Fig. P3.131 both fluids are at 20°C. If V1 = 1.7 ft/s and
nocent question. If the actual flow rate is Q = 40
losses are neglected, what should the manometer reading
m 3/h, compute (a) the head loss in ft and (b) the conh ft be?
striction diameter D that causes cavitation, assuming
that the throat divides the head loss equally and that
changing the constriction causes no additional
losses.
1 in
2
P3.130
10 ft
3 in
1
25 m
Water
10 m
2 ft
D
h
P3.134
P3.131
Mercury
P3.132 Extend the siphon analysis of Example 3.14 to account
for friction in the tube, as follows. Let the friction head
loss in the tube be correlated as 5.4(Vtube)2/(2g), which
5 cm
P3.135 The 35°C water flow of Fig. P3.135 discharges to sealevel standard atmosphere. Neglecting losses, for what
nozzle diameter D will cavitation begin to occur? To
avoid ­cavitation, should you increase or decrease D from
this critical value?
Problems 217
6 ft
P3.135
1 in
3 in
D
1
2
3
neglected, compute (a) V2 and (b) the force per unit
width of the water on the spillway.
5m
0.7 m
V1
P3.136 Air, assumed frictionless, flows through a tube, exiting
to sea-level atmosphere. Diameters at 1 and 3 are 5 cm,
while D2 = 3 cm. What mass flow of air is required to
suck water up 10 cm into section 2 of Fig. P3.136?
1
2
V2
P3.139
3
P3.140 For the water channel flow of Fig. P3.140, h1 = 1.5 m,
H = 4 m, and V1 = 3 m/s. Neglecting losses and assuming uniform flow at sections 1 and 2, find the downstream depth h2, and show that two realistic solutions are
possible.
Air
10 cm
h1
Water
V1
P3.136
P3.137 In Fig. P3.137 the piston drives water at 20°C. Neglecting losses, estimate the exit velocity V2 ft/s. If D2 is further constricted, what is the limiting possible value of
V2?
D1 = 8 in
D2 = 4 in
F = 10 lbf
pa
V2
Water
h2
H
V2
P3.140
P3.141 For the water channel flow of Fig. P3.141, h1 = 0.45 ft,
H = 2.2 ft, and V1 = 16 ft/s. Neglecting losses and assuming uniform flow at sections 1 and 2, find the downstream depth h2; show that two realistic solutions are
possible.
pa
h2
P3.137
P3.138 For the sluice gate flow of Example 3.10, use Bernoulli’s
equation, along the surface, to estimate the flow rate Q
as a function of the two water depths. Assume constant
width b.
P3.139 In the spillway flow of Fig. P3.139, the flow is assumed
uniform and hydrostatic at sections 1 and 2. If losses are
V2
h1
H
V1
P3.141
218
Chapter 3 Integral Relations for a Control Volume
*P3.142 A cylindrical tank of diameter D contains liquid to an
initial height h0. At time t = 0 a small stopper of diameter d is removed from the bottom. Using Bernoulli’s
equation with no losses, derive (a) a differential equation for the free-surface height h(t) during draining and
(b) an expression for the time t0 to drain the entire
tank.
*P3.143 The large tank of incompressible liquid in Fig. P3.143 is
at rest when, at t = 0, the valve is opened to the atmosphere. Assuming h ≈ constant (negligible velocities and
accelerations in the tank), use the unsteady frictionless
Bernoulli equation to derive and solve a differential
equation for V(t) in the pipe.
Pump
patm = 100 kPa
z
D = 3 cm
y
Gasoline,
SG = 0.68
x
P3.146
h ≈ constant
D
Valve
P3.147 The very large water tank in Fig. P3.147 is discharging
through a 4-in-diameter pipe. The pump is running, with
a performance curve hp ≈ 40 − 4 Q2, with hp in feet and
Q in ft3/s. Estimate the discharge flow rate in ft3/s if the
pipe friction loss is 1.5(V 2/2g).
V (t)
L
30 ft
P3.143
P3.144 A fire hose, with a 2-in-diameter nozzle, delivers a
water jet straight up against a ceiling 8 ft higher. The
force on the ceiling, due to momentum change, is 25
lbf. Use B
­ ernoulli’s equation to estimate the hose
flow rate, in gal/min. [Hint: The water jet area expands upward.]
P3.145 The incompressible flow form of Bernoulli’s relation,
Eq. (3.54), is accurate only for Mach numbers less
than about 0.3. At higher speeds, variable density must
be accounted for. The most common assumption for
compressible fluids is isentropic flow of an ideal gas,
or p = Cρk, where k = cp/cυ. Substitute this relation into
Eq. (3.52), integrate, and eliminate the constant C.
Compare your compressible r­ esult with Eq. (3.54) and
comment.
P3.146 The pump in Fig. P3.146 draws gasoline at 20°C from a
reservoir. Pumps are in big trouble if the liquid vaporizes
(cavitates) before it enters the pump. (a) Neglecting
losses and assuming a flow rate of 65 gal/min, find the
limitations on (x, y, z) for avoiding cavitation. (b) If pipe
friction losses are included, what additional limitations
might be ­important?
Q, V
Pump
P3.147
P3.148 By neglecting friction, (a) use the Bernoulli equation
­between surfaces 1 and 2 to estimate the volume flow
through the orifice, whose diameter is 3 cm. (b) Why is
the result to part (a) absurd? (c) Suggest a way to resolve
this paradox and find the true flow rate.
1
pa
2
4m
Water
2.5 m
1m
P3.148
Problems 219
The angular momentum theorem
θ
P3.149 The horizontal lawn sprinkler in Fig. P3.149 has a
water flow rate of 4.0 gal/min introduced vertically
through the center. Estimate (a) the retarding torque
required to keep the arms from rotating and (b)
the rotation rate (r/min) if there is no retarding
torque.
d = 7 mm
R=
d = 1–4 in
P3.149
P3.150 In Prob. P3.60 find the torque caused around flange 1 if
the center point of exit 2 is 1.2 m directly below the
flange center.
P3.151 The wye joint in Fig. P3.151 splits the pipe flow into
equal amounts Q/2, which exit, as shown, a distance
R 0 from the axis. Neglect gravity and friction.
Find an expression for the torque T about the x axis
required to keep the system rotating at angular
velocity Ω.
cm
θ
θ
P3.153
R = 6 in
15
P3.154 Water at 20°C flows at 30 gal/min through the 0.75-in-­
diameter double pipe bend of Fig. P3.154. The pressures
are p1 = 30 lbf/in2 and p2 = 24 lbf/in2. Compute the torque
T at point B necessary to keep the pipe from rotating.
B
50°
1
3 ft
2
P3.154
Q
2
T, Ω
Q
R0 >> Dpipes
θ
θ
x
P3.155 The centrifugal pump of Fig. P3.155 has a flow rate Q
and exits the impeller at an angle θ2 relative to the blades,
as shown. The fluid enters axially at section 1. Assuming
­incompressible flow at shaft angular velocity ω, derive a
formula for the power P required to drive the impeller.
R0
Q
2
Vrel, 2
θ2
P3.151
P3.152 Modify Example 3.19 so that the arm starts from rest and
spins up to its final rotation speed. The moment of inertia of the arm about O is I0. Neglecting air drag, find dω/
dt and integrate to determine the angular velocity ω(t),
assuming ω = 0 at t = 0.
P3.153 The three-arm lawn sprinkler of Fig. P3.153 receives
20°C water through the center at 2.7 m3/h. If collar friction is negligible, what is the steady rotation rate in r/min
for (a) θ = 0° and (b) θ = 40°?
R2
Blade
R1
b2
Q
T, P, ω
P3.155
P3.156 A simple turbomachine is constructed from a disk with
two internal ducts that exit tangentially through square
holes, as in Fig. P3.156. Water at 20°C enters normal to
220
Chapter 3 Integral Relations for a Control Volume
the disk at the center, as shown. The disk must drive, *P3.161 Extend Prob. P3.46 to the problem of computing the cenat 250 r/min, a small device whose retarding torque is
ter of pressure L of the normal face Fn, as in Fig. P3.161.
1.5 N · m. What is the proper mass flow of water, in kg/s?
(At the center of pressure, no moments are required to
hold the plate at rest.) Neglect friction. Express your result in terms of the sheet thickness h1 and the angle θ
2 cm
between the plate and the oncoming jet 1.
2 cm
V
h2
ρ, V
32 cm
h1
Q
L F
n
P3.156
P3.161
P3.157 Reverse the flow in Fig. P3.155, so that the system operates as a radial-inflow turbine. Assuming that the outflow into section 1 has no tangential velocity, derive an
expression for the power P extracted by the turbine.
P3.158 Revisit the turbine cascade system of Prob. P3.78, and
­derive a formula for the power P delivered, using the
­angular momentum theorem of Eq. (3.59).
P3.159 A centrifugal pump impeller delivers 4000 gal/min of
­water at 20°C with a shaft rotation rate of 1750 r/min.
­Neglect losses. If r1 = 6 in, r2 = 14 in, b1 = b2 = 1.75 in,
Vt1 = 10 ft/s, and Vt2 = 110 ft/s, compute the absolute
­velocities (a) V1 and (b) V2 and (c) the horsepower
­required. (d ) Compare with the ideal horsepower required.
P3.160 The pipe bend of Fig. P3.160 has D1 = 27 cm and D2 =
13 cm. When water at 20°C flows through the pipe at
4000 gal/min, p1 = 194 kPa (gage). Compute the torque
required at point B to hold the bend stationary.
P3.162 The waterwheel in Fig. P3.162 is being driven at 200 r/
min by a 150-ft/s jet of water at 20°C. The jet diameter is
2.5 in. Assuming no losses, what is the horsepower developed by the wheel? For what speed Ω r/min will the
horsepower developed be a maximum? Assume that
there are many buckets on the waterwheel.
Ω
4 ft
50 cm
150 ft/s
75°
C
50 cm
P3.160
V2 , p2 = pa
2
B
V1, p1
1
h3
V
P3.162
P3.163 A rotating dishwasher arm delivers at 60°C to six nozzles, as in Fig. P3.163. The total flow rate is 3.0 gal/min.
Each nozzle has a diameter of 163 in. If the nozzle flows
are equal and friction is neglected, estimate the steady
rotation rate of the arm, in r/min.
Problems 221
5 in
5 in
is 45 m wide and 2.7 m deep. If heat losses to the atmosphere and ground are negligible, estimate the downstream river conditions (Q0, T0).
6 in
Qi , Ti
40°
P3.163
T
*P3.164 A liquid of density ρ flows through a 90° bend as shown
in Fig. P3.164 and issues vertically from a uniformly
porous section of length L. Neglecting pipe and liquid
weight, ­derive an expression for the torque M at point 0
required to hold the pipe stationary.
R
y
Q
L
T + ∆T
Vw
0
Q
Power
plant
Closed
valve
x
Q0, T0
P3.166
d <<R, L
Q
P3.164
The energy equation
P3.165 There is a steady isothermal flow of water at 20°C through
the device in Fig. P3.165. Heat-transfer, gravity, and temperature effects are negligible. Known data are D1 = 9 cm,
Q1 = 220 m3/h, p1 = 150 kPa, D2 = 7 cm, Q2 = 100 m3/h,
p2 = 225 kPa, D3 = 4 cm, and p3 = 265 kPa. Compute the
rate of shaft work done for this device and its direction.
3
P3.167 For the conditions of Prob. P3.166, if the power plant is
to heat the nearby river water by no more than 12°C,
what should be the minimum flow rate Q, in m3/s,
through the plant heat exchanger? How will the value of
Q affect the downstream conditions (Q0, T0)?
P3.168 Multnomah Falls in the Columbia River Gorge has a
sheer drop of 543 ft. Using the steady flow energy equation, ­estimate the water temperature change in °F caused
by this drop.
P3.169 When the pump in Fig. P3.169 draws 220 m3/h of water at
20°C from the reservoir, the total friction head loss is 5 m.
The flow discharges through a nozzle to the atmosphere.
Estimate the pump power in kW delivered to the water.
D = 12 cm
2
2m
Isothermal
steady
flow
6m
1
Water
P3.165
P3.166 A power plant on a river, as in Fig. P3.166, must eliminate 55 MW of waste heat to the river. The river conditions ­upstream are Qi = 2.5 m3/s and Ti = 18°C. The river
P3.169
Pump
De = 5 cm
Ve
222
Chapter 3 Integral Relations for a Control Volume
P3.170 A steam turbine operates steadily under the following
­conditions. At the inlet, p = 2.5 MPa, T = 450°C, and V
= 40 m/s. At the outlet, p = 22 kPa, T = 70°C, and V =
225 m/s. (a) If we neglect elevation changes and heat
transfer, how much work is delivered to the turbine
blades, in kJ/kg? (b) If the mass flow is 10 kg/s, how
much total power is delivered? (c) Is the steam wet as it
leaves the exit?
P3.171 Consider a turbine extracting energy from a penstock in
a dam, as in Fig. P3.171. For turbulent pipe flow (Chap.
6), the friction head loss is approximately hf = CQ2,
where the constant C depends on penstock dimensions
and the properties of water. Show that, for a given penstock geometry and variable river flow Q, the maximum
turbine power possible in this case is Pmax = 2ρgHQ/3
and occurs when the flow rate is Q = √H/(3C).
planned to place pumping stations every 10 mi along the
pipe. ­Estimate the horsepower that must be delivered to
the oil by each pump.
P3.174 The pump-turbine system in Fig. P3.174 draws water
from the upper reservoir in the daytime to produce power
for a city. At night, it pumps water from lower to upper
reservoirs to restore the situation. For a design flow rate
of 15,000 gal/min in either direction, the friction head
loss is 17 ft. Estimate the power in kW (a) extracted by
the turbine and (b) delivered by the pump.
Z1 = 150 ft
1
Water at 20°C
Pumpturbine
2
Q
Z 2 = 25 ft
Penstock
H
P3.174
Turbine
P3.171
P3.172 The long pipe in Fig. P3.172 is filled with water at 20°C.
When valve A is closed, p1 − p2 = 75 kPa. When the
valve is open and water flows at 500 m3/h, p1 − p2 = 160
kPa. What is the friction head loss between 1 and 2, in m,
for the flowing condition?
2
Constantdiameter
pipe
P3.175 Water at 20°C is delivered from one reservoir to another through a long 8-cm-diameter pipe. The lower
reservoir has a surface elevation z2 = 80 m. The friction
loss in the pipe is correlated by the formula h loss ≈
17.5(V2/2g), where V is the average velocity in the pipe.
If the steady flow rate through the pipe is 500 gallons
per minute, estimate the surface elevation of the higher
reservoir.
P3.176 A fireboat draws seawater (SG = 1.025) from a submerged pipe and discharges it through a nozzle, as in
Fig. P3.176. The total head loss is 6.5 ft. If the pump
efficiency is 75 percent, what horsepower motor is required to drive it?
Pump
D = 2 in
120 ft/s
1
P3.172
10 ft
A
P3.173 A 36-in-diameter pipeline carries oil (SG = 0.89) at
1 million barrels per day (bbl/day) (1 bbl = 42 U.S. gal).
The friction head loss is 13 ft/1000 ft of pipe. It is
6 ft
P3.176
D = 6 in
Problems 223
P3.177 A device for measuring liquid viscosity is shown in Fig.
P3.177. With the parameters (ρ, L, H, d ) known, the
flow rate Q is measured and the viscosity calculated, assuming a laminar-flow pipe loss from Chap. 6, hf =
(32µLV)/(ρgd2). Heat transfer and all other losses are
negligible. (a) Derive a formula for the viscosity µ of the
fluid. (b) Calculate µ for the case d = 2 mm, ρ = 800 kg/
m3, L = 95 cm, H = 30 cm, and Q = 760 cm3/h. (c) What
is your guess of the fluid in part (b)? (d ) Verify that the
Reynolds number Red is less than 2000 (laminar pipe
flow).
saturated conditions. The mass flow is 2.5 lbm/s, and
the heat losses are 7 Btu/lb of steam. If head losses are
negligible, how much horsepower does the turbine develop?
P3.180 Water at 20°C is pumped at 1500 gal/min from the
lower to the upper reservoir, as in Fig. P3.180. Pipe
friction losses are approximated by hf ≈ 27V2/(2g),
where V is the average velocity in the pipe. If the pump
is 75 percent efficient, what horsepower is needed to
drive it?
z2 = 150 ft
Water level
z1 = 50 ft
D = 6 in
H
Pump
P3.180
P3.181 A typical pump has a head that, for a given shaft rotation
rate, varies with the flow rate, resulting in a pump performance curve as in Fig. P3.181. Suppose that this
pump is 75 percent efficient and is used for the system in
Prob. 3.180. Estimate (a) the flow rate, in gal/min, and
(b) the horsepower needed to drive the pump.
L
d
300
Q
P3.177
P3.178 The horizontal pump in Fig. P3.178 discharges 20°C water at 57 m3/h. Neglecting losses, what power in kW is
delivered to the water by the pump?
120 kPa
400 kPa
P3.178
D2 = 3 cm
Pump
Head, ft
Pump performance
200
100
0
0
1
2
Flow rate, ft3/s
3
4
P3.181
D1 = 9 cm
P3.179 Steam enters a horizontal turbine at 350 lbf/in2 absolute,
580°C, and 12 ft/s and is discharged at 110 ft/s and 25°C
P3.182 The insulated tank in Fig. P3.182 is to be filled from a
high-pressure air supply. Initial conditions in the tank are
T = 20°C and p = 200 kPa. When the valve is opened,
the initial mass flow rate into the tank is 0.013 kg/s. Assuming an ideal gas, estimate the initial rate of temperature rise of the air in the tank.
224
Chapter 3 Integral Relations for a Control Volume
Air supply:
Valve
T1 = 20°C
Tank : υ = 200 L
P1 = 1500 kPa
P3.184 The large turbine in Fig. P3.184 diverts the river flow
under a dam as shown. System friction losses are hf =
3.5V2/(2g), where V is the average velocity in the supply
pipe. For what river flow rate in m3/s will the power
­extracted be 25 MW? Which of the two possible solutions has a better “conversion efficiency”?
P3.182
z1 = 50 m
P3.183 The pump in Fig. P3.183 creates a 20°C water jet oriented to travel a maximum horizontal distance. System
friction head losses are 6.5 m. The jet may be approximated by the trajectory of frictionless particles. What
power must be ­delivered by the pump?
D=4m
Jet
15 m
De = 5 cm
D = 10 cm
25 m
P3.184
Turbine
z 2 = 10 m
z3 = 0 m
2m
Pump
P3.183
Word Problems
W3.1 Derive a control volume form of the second law of thermodynamics. Suggest some practical uses for your relation in analyzing real fluid flows.
W3.2 Suppose that it is desired to estimate volume flow Q in a
pipe by measuring the axial velocity u(r) at specific
points. For cost reasons only three measuring points are
to be used. What are the best radii selections for these
three points?
W3.3 Consider water flowing by gravity through a short pipe
connecting two reservoirs whose surface levels differ by
an amount Δz. Why does the incompressible frictionless
Bernoulli equation lead to an absurdity when the flow
rate through the pipe is computed? Does the paradox
have something to do with the length of the short pipe?
Does the paradox disappear if we round the entrance and
exit edges of the pipe?
W3.4 Use the steady flow energy equation to analyze flow
through a water faucet whose supply pressure is p0. What
physical mechanism causes the flow to vary continuously from zero to maximum as we open the faucet
valve?
W3.5 Consider a long sewer pipe, half full of water, sloping
downward at angle θ. Antoine Chézy in 1768 determined
that the average velocity of such an open-channel flow
should be V ≈ C √R tan θ, where R is the pipe radius and
C is a constant. How does this famous formula relate to
the steady flow energy equation applied to a length L of
the channel?
W3.6 Put a table tennis ball in a funnel, and attach the small
end of the funnel to an air supply. You probably won’t be
able to blow the ball either up or down out of the funnel.
­Explain why.
W3.7 How does a siphon work? Are there any limitations
(such as how high or how low can you siphon water away
from a tank)? Also, how far—could you use a flexible
tube to ­siphon water from a tank to a point 100 ft away?
Fundamentals of Engineering Exam Problems
FE3.1 In Fig. FE3.1 water exits from a nozzle into atmospheric
pressure of 101 kPa. If the flow rate is 160 gal/min, what
is the average velocity at section 1?
(a) 2.6 m/s, (b) 0.81 m/s, (c) 93 m/s, (d) 23 m/s, (e) 1.62 m/s
FE3.2 In Fig. FE3.1 water exits from a nozzle into atmospheric
pressure of 101 kPa. If the flow rate is 160 gal/min and
friction is neglected, what is the gage pressure at section 1?
(a) 1.4 kPa, (b) 32 kPa, (c) 43 kPa, (d ) 29 kPa, (e) 123 kPa
Comprehensive Problems 225
FE3.3 In Fig. FE3.1 water exits from a nozzle into atmospheric
pressure of 101 kPa. If the exit velocity is V2 = 8 m/s and
friction is neglected, what is the axial flange force required to keep the nozzle attached to pipe 1?
(a) 11 N, (b) 56 N, (c) 83 N, (d ) 123 N, (e) 110 N
(2)
d = 4 cm
patm
70 cm
d = 12 cm
(1)
7 cm
4 cm
(1)
(2)
Jet
Pump
patm = 101 kPa
Water
h
FE3.6
FE3.1
FE3.4 In Fig. FE3.1 water exits from a nozzle into atmospheric
pressure of 101 kPa. If the manometer fluid has a specific gravity of 1.6 and h = 66 cm, with friction neglected, what is the average velocity at section 2?
(a) 4.55 m/s, (b) 2.4 m/s, (c) 2.95 m/s, (d ) 5.55 m/s, (e) 3.4 m/s
FE3.5 A jet of water 3 cm in diameter strikes normal to a plate
as in Fig. FE3.5. If the force required to hold the plate is
23 N, what is the jet velocity?
(a) 2.85 m/s, (b) 5.7 m/s, (c) 8.1 m/s, (d ) 4.0 m/s, (e) 23 m/s
3 cm
V
120 cm
F = 23 N
FE3.5
FE3.6 A fireboat pump delivers water to a vertical nozzle with
a 3:1 diameter ratio, as in Fig. FE3.6. If friction is neglected and the flow rate is 500 gal/min, how high will
the outlet water jet rise?
(a) 2.0 m, (b) 9.8 m, (c) 32 m, (d ) 64 m, (e) 98 m
FE3.7 A fireboat pump delivers water to a vertical nozzle with
a 3:1 diameter ratio, as in Fig. FE3.6. If friction is neglected and the pump increases the pressure at section 1
to 51 kPa (gage), what will be the resulting flow rate?
(a) 187 gal/min, (b) 199 gal/min, (c) 214 gal/min,
(d) 359 gal/min, (e) 141 gal/min
FE3.8 A fireboat pump delivers water to a vertical nozzle with
a 3:1 diameter ratio, as in Fig. FE3.6. If duct and nozzle
friction are neglected and the pump provides 12.3 ft of
head to the flow, what will be the outlet flow rate?
(a) 85 gal/min, (b) 120 gal/min, (c) 154 gal/min,
(d) 217 gal/min, (e) 285 gal/min
FE3.9 Water flowing in a smooth 6-cm-diameter pipe enters
a venturi contraction with a throat diameter of 3 cm.
­Upstream pressure is 120 kPa. If cavitation occurs in the
throat at a flow rate of 155 gal/min, what is the estimated
fluid vapor pressure, assuming ideal frictionless flow?
(a) 6 kPa, (b) 12 kPa, (c) 24 kPa, (d) 31 kPa, (e) 52 kPa
FE3.10 Water flowing in a smooth 6-cm-diameter pipe enters a
venturi contraction with a throat diameter of 4 cm.
­Upstream pressure is 120 kPa. If the pressure in the
throat is 50 kPa, what is the flow rate, assuming ideal
frictionless flow?
(a) 7.5 gal/min, (b) 236 gal/min, (c) 263 gal/min,
(d) 745 gal/min, (e) 1053 gal/min
Comprehensive Problems
C3.1
In a certain industrial process, oil of density ρ flows
through the inclined pipe in Fig. C3.1. A U-tube
manometer, with fluid density ρm, measures the pressure
difference between points 1 and 2, as shown. The pipe
226
Chapter 3 Integral Relations for a Control Volume
flow is steady, so that the fluids in the manometer are
stationary. (a) Find an ­analytic expression for p1 − p2 in
terms of the system p­ arameters. (b) Discuss the conditions on h necessary for there to be no flow in the pipe.
(c) What about flow up, from 1 to 2? (d ) What about
flow down, from 2 to 1?
Ve
R
h
Vj
F
(2)
θ
C3.3
(1)
s
ρ
h
ρm
C3.1
C3.2
C3.3
L
A rigid tank of volume 𝒱 = 1.0 m3 is initially filled with
air at 20°C and p0 = 100 kPa. At time t = 0, a vacuum
pump is turned on and evacuates air at a constant volume flow rate Q = 80 L/min (regardless of the pressure). Assume an ideal gas and an isothermal process.
(a) Set up a differential equation for this flow. (b) Solve
this equation for t as a function of (𝒱, Q, p, p0). (c)
Compute the time in minutes to pump the tank down to
p = 20 kPa. Hint: Your answer should lie between 15
and 25 min.
Suppose the same steady water jet as in Prob. P3.40 (jet
velocity 8 m/s and jet diameter 10 cm) impinges instead on a cup cavity as shown in Fig. C3.3. The water
is turned 180° and exits, due to friction, at lower velocity, Ve = 4 m/s. (Looking from the left, the exit jet is a
circular ­annulus of outer radius R and thickness h,
flowing toward the viewer.) The cup has a radius of
curvature of 25 cm. Find (a) the thickness h of the exit
jet and (b) the force F required to hold the cupped object in place. (c) Compare part (b) to Prob. 3.40, where
F ≈ 500 N, and give a physical explanation as to why F
has changed.
C3.4
Ve
The airflow underneath an air hockey puck is very complex, especially since the air jets from the air hockey table impinge on the underside of the puck at various
points nonsymmetrically. A reasonable approximation is
that at any given time, the gage pressure on the bottom of
the puck is halfway between zero (atmospheric pressure)
and the stagnation pressure of the impinging jets. (Stagnation pressure is defined as p0 = 12 ρV2jet .) (a) Find the
jet velocity Vjet ­required to support an air hockey puck of
weight W and diameter d. Give your answer in terms of
W, d, and the density ρ of the air. (b) For W = 0.05 lbf
and d = 2.5 in, estimate the required jet velocity in ft/s.
z2
V
Atmosphere
Fan
C3.5
z1
References 227
C3.5
Neglecting friction sometimes leads to odd results. You
are asked to analyze and discuss the following example
in Fig. C3.5. A fan blows air through a duct from section
1 to section 2, as shown. Assume constant air density ρ.
­ eglecting frictional losses, find a relation between the
N
­required fan head hp and the flow rate and the elevation change. Then explain what may be an unexpected
­result.
Design Project
D3.1
where hp is the pump head (ft), n is the shaft rotation rate
(r/s), and Dp is the impeller diameter (ft). The range of
­validity is 0 < ζ < 0.027. The pump of Fig. P3.181 had
Dp = 2 ft in diameter and rotated at n = 20 r/s (1200 r/
min). The solution to Prob. P3.181, namely, Q ≈ 2.57
ft3/s and hp ≈ 172 ft, corresponds to ϕ ≈ 3.46, ζ ≈ 0.016,
η ≈ 0.75 (or 75 percent), and power to the water = ρgQhp
≈ 27,500 ft · lbf/s (50 hp). Please check these numerical
­values before beginning this project.
Now revisit Prob. P3.181 and select a low-cost pump
that rotates at a rate no slower than 600 r/min and delivers no less than 1.0 ft3/s of water. Assume that the cost of
the pump is linearly proportional to the power input required. Comment on any limitations to your results.
Let us generalize Probs. P3.180 and P3.181, in which a
pump performance curve was used to determine the flow
rate between reservoirs. The particular pump in Fig.
P3.181 is one of a family of pumps of similar shape,
whose dimensionless performance is as follows:
Head:
ϕ ≈ 6.04 − 161ζ
ϕ=
gh
2
n
D2p
and ζ =
Q
nD3p
Efficiency:
η ≈ 70ζ − 91,500ζ3
η=
power to water
power input
References
1.
2.
3.
4.
5.
6.
D. T. Greenwood, Advanced Dynamics, Cambridge
­University Press, New York, 2006.
T. von Kármán, The Wind and Beyond, Little, Brown,
­Boston, 1967.
J. P. Holman, Heat Transfer, 10th ed., McGraw-Hill,
New York, 2009.
A. G. Hansen, Fluid Mechanics, Wiley, New York, 1967.
M. C. Potter, D. C. Wiggert, and M. Hondzo, Mechanics of
Fluids, Brooks/Cole, Chicago, 2001.
S. Klein and G. Nellis, Thermodynamics, Cambridge
­University Press, New York, 2011.
7.
8.
9.
10.
11.
Y. A. Cengel and M. A. Boles, Thermodynamics: An
­Engineering Approach, 7th ed., McGraw-Hill, New York,
2010.
J. F. Wendt, Computational Fluid Dynamics: An
­Introduction, Springer, 3d ed., New York, 2009.
W. G. Vincenti, “Control Volume Analysis: A Difference
in Thinking between Engineering and Physics,” Technology and Culture, vol. 23, no. 2, 1982, pp. 145–174.
J. Keenan, Thermodynamics, Wiley, New York, 1941.
J. Hunsaker and B. Rightmire, Engineering Applications of
Fluid Mechanics, McGraw-Hill, New York, 1947.
The differential equations to be studied in this chapter can be modeled numerically
by computational fluid dynamics (CFD). This figure shows a NASA computergenerated model of the Space Shuttle during reentry in 1991. CFD has supplanted
wind tunnels for many evaluations of aircraft. The computation processing speed in
NASA increased from 0.2 gigaflops on Cray X-MP supercomputer in the early 1980s
to 5.95 petaflops on Pleiades in 2019. As computing power increases and computer
models become more sophisticated, CFD has become a powerful tool for fluid
mechanics and aeronautics research and industries.
Source: The NASA Library/Alamy Stock Photo.
228
Chapter 4
Differential Relations
for Fluid Flow
Motivation. In analyzing fluid motion, we might seek the point-by-point details
of a flow pattern by analyzing an infinitesimal region of the flow. This is the
“differential” approach and is the subject of this chapter.
This chapter treats the second in our trio of techniques for analyzing fluids.
That is, we apply our four basic conservation laws to an infinitesimally small
control volume or, alternately, to an infinitesimal fluid system. In either case the
results yield the basic differential equations of fluid motion. Appropriate boundary conditions are also developed.
In their most basic form, these differential equations of motion are quite
difficult to solve, and very little is known about their general mathematical
properties. However, certain things can be done that have great educational
value. First, as shown in Chap. 5, the equations (even if unsolved) reveal the
basic dimensionless parameters that govern fluid motion. Second, as shown in
Chap. 6, a great number of useful solutions can be found if one makes two
simplifying assumptions: (1) steady flow and (2) incompressible flow. A third
and rather drastic simplification, frictionless flow, makes our old friend the
Bernoulli equation valid and yields a wide variety of idealized, or perfect-fluid,
possible solutions. These idealized flows are treated in Chap. 8, and we must
be careful to ascertain whether such solutions are in fact realistic when compared with actual fluid motion. Finally, even the difficult general differential
equations now yield to the approximating technique known as computational
fluid dynamics (CFD) whereby the derivatives are simulated by algebraic relations between a finite number of grid points in the flow field, which are then
solved on a computer. Reference 1 is an example of a textbook devoted entirely
to numerical analysis of fluid motion.
229
230
Chapter 4 Differential Relations for Fluid Flow
4.1 The Acceleration Field of a Fluid
In Sec. 1.7 we established the cartesian vector form of a velocity field that varies
in space and time:
V(r, t) = iu(x, y, z, t) + jυ(x, y, z, t) + kw(x, y, z, t)
(1.4)
This is the most important variable in fluid mechanics: Knowledge of the velocity
vector field is nearly equivalent to solving a fluid flow problem. Our coordinates are
fixed in space, and we observe the fluid as it passes by—as if we had scribed a set
of coordinate lines on a glass window in a wind tunnel. This is the Eulerian frame of
reference, as opposed to the Lagrangian frame, which follows the moving position of
individual particles.
To write Newton’s second law for an infinitesimal fluid system, we need to
calculate the acceleration vector field a of the flow. Thus, we compute the total
time derivative of the velocity vector:
a=
dV
du
dυ
dw
=i
+j
+k
dt
dt
dt
dt
Since each scalar component (u, υ, w) is a function of the four variables (x, y, z,
t), we use the chain rule to obtain each scalar time derivative. For example,
du(x, y, z, t) ∂u ∂u dx ∂u dy ∂u dz
=
+
+
+
dt
∂t
∂x dt
∂y dt
∂z dt
But, by definition, dx/dt is the local velocity component u, and dy/dt = υ, and
dz/dt = w. The total time derivative of u may thus be written as follows, with
exactly similar expressions for the time derivatives of υ and w:
ax =
du ∂u
∂u
∂u
∂u ∂u
=
+u
+υ
+w
=
+ (V · ∇ )u
dt
∂t
∂x
∂y
∂z
∂t
ay =
dυ ∂υ
∂υ
∂υ
∂υ ∂υ
=
+u
+υ
+w
=
+ (V · ∇ )υ
dt
∂t
∂x
∂y
∂z
∂t
az =
dw ∂w
∂w
∂w
∂w
∂w
=
+u
+υ
+w
=
+ (V · ∇ )w
dt
∂t
∂x
∂y
∂z
∂t
(4.1)
Summing these into a vector, we obtain the total acceleration:
a=
dV ∂V
∂V
∂V
∂V
∂V
=
+ (u
+υ
+w
=
+ (V · ∇)V (4.2)
dt
∂t
∂x
∂y
∂z )
∂t
Local
Convective
The term 𝜕V/𝜕t is called the local acceleration, which vanishes if the flow is
steady—that is, independent of time. The three terms in parentheses are called
the convective acceleration, which arises when the particle moves through regions
of spatially varying velocity, as in a nozzle or diffuser. Flows that are nominally
“steady” may have large accelerations due to the convective terms.
4.1 The Acceleration Field of a Fluid 231
Note our use of the compact dot product involving V and the gradient
operator ∇:
u
∂
∂
∂
∂
∂
∂
+j
+k
+υ
+ w = V · ∇ where ∇ = i
∂x
∂y
∂z
∂x
∂y
∂z
The total time derivative—sometimes called the substantial or material derivative—concept may be applied to any variable, such as the pressure:
dp ∂p
∂p
∂p
∂p ∂p
=
+u
+υ
+w
=
+ (V · ∇)p(4.3)
dt
∂t
∂x
∂y
∂z
∂t
Wherever convective effects occur in the basic laws involving mass, momentum,
or energy, the basic differential equations become nonlinear and are usually more
complicated than flows that do not involve convective changes.
We emphasize that this total time derivative follows a particle of fixed identity,
making it convenient for expressing laws of particle mechanics in the eulerian
fluid field description. The operator d/dt is sometimes assigned a special symbol
such as D/Dt as a further reminder that it contains four terms and follows a fixed
particle.
As another reminder of the special nature of d/dt, some writers give it the
name substantial or material derivative.
EXAMPLE 4.1
Given the Eulerian velocity vector field
V = 3ti + xzj + ty2k
find the total acceleration of a particle.
Solution
∙ Assumptions: Given three known unsteady velocity components, u = 3t, υ = xz, and
w = ty2.
∙ Approach: Carry out all the required derivatives with respect to (x, y, z, t), substitute into the total acceleration vector, Eq. (4.2), and collect terms.
∙ Solution step 1: First work out the local acceleration 𝜕V/𝜕t:
∂V
∂u
∂υ
∂w
∂
∂
∂
=i
+j
+k
= i (3t) + j (xz) + k (ty2 ) = 3i + 0j + y2 k
∂t
∂t
∂t
∂t
∂t
∂t
∂t
∙ Solution step 2:
Eq. (4.2), are
∂V
=(3t)
∂x
∂V
υ
=(xz)
∂y
u
w
In a similar manner, the convective acceleration terms, from
∂
(3ti + xzj + ty2k) = (3t) (0i + zj + 0k) = 3tz j
∂x
∂
(3ti + xzj + ty2k) = (xz) (0i + 0j + 2tyk) = 2txyz k
∂y
∂V
∂
=(ty2 )
(3ti + xzj + ty2k) = (ty2 ) (0i + xj + 0k) = txy2 j
∂z
∂z
232
Chapter 4 Differential Relations for Fluid Flow
∙ Solution step 3:
tial” derivative:
Combine all four terms above into the single “total” or “substan-
d V ∂V
∂V
∂V
∂V
=
+u
+υ
+w
= (3i + y2k) + 3tzj + 2txyzk + txy2j
dt
∂t
∂x
∂y
∂z
= 3i + (3tz + txy2 )j + ( y2 + 2txyz)k Ans.
∙ Comments: Assuming that V is valid everywhere as given, this total acceleration
vector dV/dt applies to all positions and times within the flow field.
4.2 The Differential Equation of Mass Conservation
Conservation of mass, often called the continuity relation, states that the fluid
mass cannot change. We apply this concept to a very small region. All the basic
differential equations can be derived by considering either an elemental control
volume or an elemental system. We choose an infinitesimal fixed control volume
(dx, dy, dz), as in Fig. 4.1, and use our basic control volume relations from
Chap. 3. The flow through each side of the element is approximately onedimensional, and so the appropriate mass conservation relation to use here is
∂ρ
d 𝒱 + ∑ (ρi Ai Vi ) out − ∑ (ρi AiVi ) in = 0
(3.22)
∂t
i
i
CV
∫
The element is so small that the volume integral simply reduces to a differential term:
∫
CV
∂ρ
∂ρ
d𝒱 ≈
dx dy dz
∂t
∂t
The mass flow terms occur on all six faces, three inlets and three outlets. We make
use of the field or continuum concept from Chap. 1, where all fluid properties are
considered to be uniformly varying functions of time and position, such as ρ =
ρ(x, y, z, t). Thus, if T is the temperature on the left face of the element in Fig. 4.1,
the right face will have a slightly different temperature T + (∂T/∂x) dx. For mass
conservation, if ρu is known on the left face, the value of this product on the right
face is ρu + (∂ρu/∂x) dx.
y
Control volume
ρ u + ∂ ( ρ u) dx dy dz
ρ u dy dz
∂x
dy
x
dz
Fig. 4.1 Elemental cartesian fixed
control volume showing the inlet
and outlet mass flows on the x faces.
z
dx
4.2 The Differential Equation of Mass Conservation 233
Figure 4.1 shows only the mass flows on the x or left and right faces. The
flows on the y (bottom and top) and the z (back and front) faces have been omitted to avoid cluttering up the drawing. We can list all these six flows as follows:
Face
Inlet mass flow
Outlet mass flow
x
ρu dy dz
∂
[ ρu + ∂x (ρu) dx ]dy dz
y
ρυ dx dz
z
ρw dx dy
∂
[ ρυ + ∂y (ρυ) dy ] dx dz
∂
[ ρw + ∂z (ρw) dz ] dx dy
Introduce these terms into Eq. (3.22) and we have
∂ρ
∂
∂
∂
dx dy dz +
(ρu) dx dy dz +
(ρυ) dx dy dz +
(ρw) dx dy dz = 0
∂t
∂x
∂y
∂z
The element volume cancels out of all terms, leaving a partial differential equation involving the derivatives of density and velocity:
∂ρ
∂
∂
∂
+
(ρu) +
(ρυ) +
(ρw) = 0
∂t
∂x
∂y
∂z
(4.4)
This is the desired result: conservation of mass for an infinitesimal control volume. It is often called the equation of continuity because it requires no assumptions except that the density and velocity are continuum functions. That is, the
flow may be either steady or unsteady, viscous or frictionless, compressible or
incompressible.1 However, the equation does not allow for any source or sink
singularities within the element.
The vector gradient operator
∂
∂
∂
+j
+k
∂x
∂y
∂z
enables us to rewrite the equation of continuity in a compact form, not that it
helps much in finding a solution. The last three terms of Eq. (4.4) are equivalent
to the divergence of the vector ρV
∇=i
∂
∂
∂
(ρu) +
(ρυ) +
(ρw) ≡ ∇ · ( ρV)
∂x
∂y
∂z
(4.5)
so the compact form of the continuity relation is
∂ρ
+ ∇ · (ρV) = 0
(4.6)
∂t
In this vector form the equation is still quite general and can readily be converted
to other coordinate systems.
1
One case where Eq. (4.4) might need special care is two-phase flow, where the density is
­discontinuous between the phases. For further details on this case, see Ref. 2, for example.
234
Chapter 4 Differential Relations for Fluid Flow
υr
υθ
r
θ
Base
line
Typical point (r, θ , z)
Typical
infinitesimal
element
υz
dr
dz
r dθ
Cyl
r dθ
ind
rica
l ax
is
Fig. 4.2 Definition sketch for the
cylindrical coordinate system.
z
Cylindrical Polar Coordinates
The most common alternative to the cartesian system is the cylindrical polarcoordinate system, sketched in Fig. 4.2. An arbitrary point P is defined by a
distance z along the axis, a radial distance r from the axis, and a rotation angle
θ about the axis. The three independent orthogonal velocity components are an
axial velocity υz, a radial velocity υr, and a circumferential velocity υθ, which is
positive counterclockwise—that is, in the direction of increasing θ. In general, all
components, as well as pressure and density and other fluid properties, are continuous functions of r, θ, z, and t.
The divergence of any vector function A(r, θ, z, t) is found by making the
transformation of coordinates
y
r = (x2 + y2 ) 1/2
θ = tan −1
z = z
(4.7)
x
2
and the result is given here without proof
1 ∂
1 ∂
∂
∇·A=
(rAr ) +
(A ) +
(Az )
(4.8)
r ∂r
r ∂θ θ
∂z
The general continuity equation (4.6) in cylindrical polar coordinates is thus
∂ρ 1 ∂
1 ∂
∂
+
(rρυr ) +
(ρυθ ) +
(ρυz ) = 0
r ∂r
r ∂θ
∂t
∂z
(4.9)
There are other orthogonal curvilinear coordinate systems, notably spherical
polar coordinates, which occasionally merit use in a fluid mechanics problem.
We shall not treat these systems here except in Prob. P4.12.
There are also other ways to derive the basic continuity equation (4.6) that are
interesting and instructive. One example is the use of the divergence theorem.
Ask your instructor about these alternative approaches.
2
See, for example, Ref. 3.
4.2 The Differential Equation of Mass Conservation 235
Steady Compressible Flow
If the flow is steady, 𝜕/𝜕t ≡ 0 and all properties are functions of position only.
Equation (4.6) reduces to
∂
∂
∂
Cartesian:
(ρu) +
(ρυ) +
(ρw) = 0
∂x
∂y
∂z
Cylindrical:
1 ∂
1 ∂
∂
(rρυr ) +
(ρυθ ) +
(ρυz ) = 0
r ∂r
r ∂θ
∂z
(4.10)
Since density and velocity are both variables, these are still nonlinear and rather
formidable, but a number of special-case solutions have been found.
Incompressible Flow
A special case that affords great simplification is incompressible flow, where the
density changes are negligible. Then 𝜕ρ/𝜕t ≈ 0 regardless of whether the flow is
steady or unsteady, and the density can be slipped out of the divergence in Eq.
(4.6) and divided out. The result
∇ · V = 0
(4.11)
is valid for steady or unsteady incompressible flow. The two coordinate forms are
Cartesian:
Cylindrical:
∂u ∂υ ∂w
+
+
= 0
∂x ∂y
∂z
(4.12a)
1 ∂
1 ∂
∂
(rυr ) +
(υ ) +
(υz ) = 0
r ∂r
r ∂θ θ
∂z
(4.12b)
These are linear differential equations, and a wide variety of solutions are known,
as discussed in Chaps. 6 to 8. Since no author or instructor can resist a wide
variety of solutions, it follows that a great deal of time is spent studying incompressible flows. Fortunately, this is exactly what should be done, because most
practical engineering flows are approximately incompressible, the chief exception
being the high-speed gas flows treated in Chap. 9.
When is a given flow approximately incompressible? We can derive a nice
criterion by using some density approximations. In essence, we wish to slip the
density out of the divergence in Eq. (4.6) and approximate a typical term such as
∂
∂u
(ρu) ≈ ρ ∂x
∂x
This is equivalent to the strong inequality
(4.13)
∣ ∣ ∣ ∣
∣ ∣ ∣ ∣
u
∂ρ
∂x
≪ ρ
∂u
∂x
δρ
δV
≪
(4.14)
ρ
V
As shown in Eq. (1.38), the pressure change is approximately proportional to the
density change and the square of the speed of sound a of the fluid:
or
δp ≈ a2 δρ
(4.15)
236
Chapter 4 Differential Relations for Fluid Flow
Meanwhile, if elevation changes are negligible, the pressure is related to the
velocity change by Bernoulli’s equation (3.52):
δp ≈ −ρV δV
(4.16)
Combining Eqs. (4.14) to (4.16), we obtain an explicit criterion for incompressible flow:
V2
= Ma2 ≪ 1
(4.17)
a2
where Ma = V/a is the dimensionless Mach number of the flow. How small is
small? The commonly accepted limit is
Ma ≤ 0.3
(4.18)
For air at standard conditions, a flow can thus be considered incompressible if
the velocity is less than about 100 m/s (330 ft/s). This encompasses a wide variety of airflows: automobile and train motions, light aircraft, landing and takeoff
of high-speed aircraft, most pipe flows, and turbomachinery at moderate rotational speeds. Further, it is clear that almost all liquid flows are incompressible,
since flow velocities are small and the speed of sound is very large.3
Before attempting to analyze the continuity equation, we shall proceed with
the derivation of the momentum and energy equations, so that we can analyze
them as a group. A very clever device called the stream function can often make
short work of the continuity equation, but we shall save it until Sec. 4.7.
One further remark is appropriate: The continuity equation is always important
and must always be satisfied for a rational analysis of a flow pattern. Any newly
discovered momentum or energy “solution” will ultimately fail when subjected
to critical analysis if it does not also satisfy the continuity equation.
EXAMPLE 4.2
Under what conditions does the velocity field
V = (a1x + b1y + c1z)i + (a2x + b2y + c2z)j + (a3x + b3y + c3z)k
where a1, b1, etc. = const, represent an incompressible flow that conserves mass?
Solution
Recalling that V = ui + υj + wk, we see that u = (a1x + b1y + c1z), etc. Substituting
into Eq. (4.12a) for incompressible continuity, we obtain
∂
∂
∂
(a1x + b1y + c1z) +
(a2x + b2y + c2z) +
(a3x + b3y + c3z) = 0
∂x
∂y
∂z
or
a1 + b2 + c3 = 0
Ans.
At least two of constants a1, b2, and c3 must have opposite signs. Continuity imposes no
restrictions whatever on constants b1, c1, a2, c2, a3, and b3, which do not contribute to a
volume increase or decrease of a differential element.
3
An exception occurs in geophysical flows, where a density change is imposed thermally or
mechanically rather than by the flow conditions themselves. An example is fresh water layered
upon saltwater or warm air layered upon cold air in the atmosphere. We say that the fluid is stratified, and we must account for vertical density changes in Eq. (4.6) even if the velocities are small.
4.2 The Differential Equation of Mass Conservation 237
EXAMPLE 4.3
An incompressible velocity field is given by
u = a(x2 − y2 )
υ unknown
w=b
where a and b are constants. What must the form of the velocity component υ be?
Solution
Again Eq. (4.12a) applies:
∂
∂υ ∂b
(ax2 − ay2 ) +
+
=0
∂x
∂y
∂z
∂υ
= −2ax
∂y
or
(1)
This is easily integrated partially with respect to y:
υ (x, y, z, t) = −2axy + f (x, z, t)
Ans.
This is the only possible form for υ that satisfies the incompressible continuity equation.
The function of integration f is entirely arbitrary since it vanishes when υ is differentiated
with respect to y.4
EXAMPLE 4.4
A centrifugal impeller of 40-cm diameter is used to pump hydrogen at 15°C and 1-atm
pressure. Estimate the maximum allowable impeller rotational speed to avoid compressibility effects at the blade tips.
Solution
∙ Assumptions: The maximum fluid velocity is approximately equal to the impeller
tip speed:
Vmax ≈ Ωrmax where rmax = D/2 = 0.20 m
∙ Approach: Find the speed of sound of hydrogen and make sure that Vmax is much less.
∙ Property values: From Table A.4 for hydrogen, R = 4124 m2/(s2 – K) and k = 1.41.
From Eq. (1.39) at 15°C = 288 K, compute the speed of sound:
aH2 = √kRT = √1.41[4124 m2/(s2 − K) ] (288 K) ≈ 1294 m/s
∙ Final solution step: Use our rule of thumb, Eq. (4.18), to estimate the maximum
impeller speed:
V = Ωrmax ≤ 0.3a
or
Ω(0.2 m) ≤ 0.3(1294 m/s)
rad
rev
Solve for
Ω ≤ 1940
≈ 18,500
Ans.
s
min
∙ Comments: This is a high rate because the speed of sound of hydrogen, a light gas,
is nearly four times greater than that of air. An impeller moving at this speed in air
would create tip shock waves.
4
This is a very realistic flow that simulates the turning of an inviscid fluid through a 60° angle;
see Examples 4.7 and 4.9.
238
Chapter 4 Differential Relations for Fluid Flow
4.3 The Differential Equation of Linear Momentum
This section uses an elemental volume to derive Newton’s law for a moving fluid.
An alternate approach, which the reader might pursue, would be a force balance on
an elemental moving particle. Having done it once in Sec. 4.2 for mass conservation,
we can move along a little faster this time. We use the same elemental control volume
as in Fig. 4.1, for which the appropriate form of the linear momentum relation is
∑F =
∂
∂t (
∫
CV
Vρ d 𝒱) +
∑ (m iVi ) out − ∑ (m iVi ) in (3.40)
Again the element is so small that the volume integral simply reduces to a derivative term:
∂
∂
(Vρ d 𝒱) ≈
(ρV) dx dy dz
(4.19)
∂t
∂t
The momentum fluxes occur on all six faces, three inlets and three outlets.
Referring again to Fig. 4.1, we can form a table of momentum fluxes by exact
analogy with the discussion that led up to the equation for net mass flux:
Faces
Inlet momentum flux
x
ρuV dy dz
y
ρυV dx dz
z
ρwV dx dy
Outlet momentum flux
∂
[ ρuV + ∂x (ρuV) dx ] dy dz
∂
[ ρυV + ∂y (ρυV) dy ] dx dz
∂
[ ρwV + ∂z (ρwV) dz ] dx dy
Introduce these terms and Eq. (4.19) into Eq. (3.40), and get this intermediate
result:
∂
∂
∂
∂
∑ F = dx dy dz[ (ρV) + (ρuV) + (ρυV) + (ρwV) ]
(4.20)
∂t
∂x
∂y
∂z
Note that this is a vector relation. A simplification occurs if we split up the term
in brackets as follows:
∂
∂
∂
∂
(ρV) +
(ρuV) +
(ρυV) +
(ρwV)
∂t
∂x
∂y
∂z
∂ρ
∂V
∂V
∂V
∂V
= V[
+ ∇ · (ρV) ] + ρ (
+u
+υ
+w
∂t
∂t
∂x
∂y
∂z )
(4.21)
The term in brackets on the right-hand side is seen to be the equation of continuity, Eq. (4.6), which vanishes identically. The long term in parentheses on the
right-hand side is seen from Eq. (4.2) to be the total acceleration of a particle
that instantaneously occupies the control volume:
∂V
∂V
∂V
∂V dV
+u
+υ
+w
=
(4.2)
∂t
∂x
∂y
∂z
dt
Thus, we have now reduced Eq. (4.20) to
∑F = ρ
dV
dx dy dz
dt
(4.22)
4.3 The Differential Equation of Linear Momentum 239
It might be good for you to stop and rest now and think about what we have just
done. What is the relation between Eqs. (4.22) and (3.40) for an infinitesimal
control volume? Could we have begun the analysis at Eq. (4.22)?
Equation (4.22) points out that the net force on the control volume must be of
differential size and proportional to the element volume. These forces are of two
types, body forces and surface forces. Body forces are due to external fields
(gravity, magnetism, electric potential) that act on the entire mass within the element. The only body force we shall consider in this book is gravity. The gravity
force on the differential mass ρ dx dy dz within the control volume is
dFgrav = ρg dx dy dz
(4.23)
where g may in general have an arbitrary orientation with respect to the coordinate system. In many applications, such as Bernoulli’s equation, we take z “up,”
and g = –gk.
The surface forces are due to the stresses on the sides of the control surface.
These stresses are the sum of hydrostatic pressure plus viscous stresses τij that
arise from motion with velocity gradients:
−p + τxx
σij = ∞ τxy
τxz
τyx
−p + τyy
τyz
τzx
τzy
∞
−p + τzz
(4.24)
The subscript notation for stresses is given in Fig. 4.3. Unlike velocity V, which is
a three-component vector, stresses σij and τij and strain rates εij are nine-component
tensors and require two subscripts to define each component. For further study of
tensor analysis, see Ref. 6, 11, or 13.
It is not these stresses but their gradients, or differences, that cause a net force
on the differential control surface. This is seen by referring to Fig. 4.4, which
y
σyy
σyx
σyz
σxy
σzy
σxx
σzx
σzz
z
Fig. 4.3 Notation for stresses.
σxz
x
σi j = Stress in j
direction on a face
normal to i axis
240
Chapter 4 Differential Relations for Fluid Flow
(σyx +
y
∂σyx
∂y
dy) dx dz
σzx dx dy
σxx dy dz
(σx x +
dy
σyx dx dz
Fig. 4.4 Elemental cartesian fixed
control volume showing the surface
forces in the x direction only.
∂σx x
dx) dy dz
∂x
x
dz
dx
z
(σzx +
∂σzx
∂z
dz) dx dy
shows only the x-directed stresses to avoid cluttering up the drawing. For example, the leftward force σxx dy dz on the left face is balanced by the rightward force
σxx dy dz on the right face, leaving only the net rightward force (∂σxx /∂x) dx dy
dz on the right face. The same thing happens on the other four faces, so the net
surface force in the x direction is given by
dFx,surf = [
∂
∂
∂
(σxx ) +
(σyx ) +
(σzx ) ] dx dy dz
∂x
∂y
∂z
(4.25)
We see that this force is proportional to the element volume. Notice that the stress
terms are taken from the top row of the array in Eq. (4.24). Splitting this row
into pressure plus viscous stresses, we can rewrite Eq. (4.25) as
∂p
dFx
∂
∂
∂
=− +
(τxx ) +
(τyx ) +
(τzx )
d𝒱
∂x ∂x
∂y
∂z
(4.26)
where d 𝒱 = dx dy dz. In an exactly similar manner, we can derive the y and z
forces per unit volume on the control surface:
dFy
d𝒱
=−
∂p
∂
∂
∂
+
(τxy ) +
(τyy ) +
(τzy )
∂y ∂x
∂y
∂z
dFz
∂p
∂
∂
∂
=− +
(τxz ) +
(τyz ) +
(τzz )
d𝒱
∂z ∂x
∂y
∂z
(4.27)
Now we multiply Eqs. (4.26) and (4.27) by i, j, and k, respectively, and add to
obtain an expression for the net vector surface force:
dF
dF
( d 𝒱 )surf = −∇p + ( d 𝒱 )viscous
(4.28)
4.3 The Differential Equation of Linear Momentum 241
where the viscous force has a total of nine terms:
∂τxx ∂τyx ∂τzx
dF
=
i
( d 𝒱 )viscous
( ∂x + ∂y + ∂z )
∂τxy ∂τyy ∂τzy
+ j(
+
+
∂x
∂y
∂z )
∂τxz ∂τyz ∂τzz
+ k(
+
+
∂x
∂y
∂z )
(4.29)
Since each term in parentheses in Eq. (4.29) represents the divergence of a stress
component vector acting on the x, y, and z faces, respectively, Eq. (4.29) is sometimes expressed in divergence form:
where
dF
( d 𝒱 )viscous = ∇ · τij
(4.30)
τxx
τij = £ τxy
τxz
(4.31)
τyx
τyy
τyz
τzx
τzy § τzz
is the viscous stress tensor acting on the element. The surface force is thus the
sum of the pressure gradient vector and the divergence of the viscous stress tensor. ­Substituting into Eq. (4.22) and utilizing Eq. (4.23), we have the basic differential momentum equation for an infinitesimal element:
ρg − ∇p + ∇ · τij = ρ
dV
dt
dV ∂V
∂V
∂V
∂V
=
+u
+υ
+w
dt
∂t
∂x
∂y
∂z
where
(4.32)
(4.33)
We can also express Eq. (4.32) in words:
Gravity force per unit volume + pressure force per unit volume
+ viscous force per unit volume = density × acceleration
(4.34)
Equation (4.32) is so brief and compact that its inherent complexity is almost
i­nvisible. It is a vector equation, each of whose component equations contains
nine terms. Let us therefore write out the component equations in full to illustrate
the mathematical difficulties inherent in the momentum equation:
∂p ∂τxx ∂τyx ∂τzx
∂u
∂u
∂u
∂u
+
+
+
=ρ( +u
+υ
+w )
∂x
∂x
∂y
∂z
∂t
∂x
∂y
∂z
ρgx −
ρgy −
ρgz −
∂p ∂τxy ∂τyy ∂τzy
∂υ
∂υ
∂υ
∂υ
+
+
+
=ρ( +u
+υ
+w )
∂y
∂x
∂y
∂z
∂t
∂x
∂y
∂z
∂p ∂τxz ∂τyz ∂τzz
∂w
∂w
∂w
∂w
+
+
+
=ρ(
+u
+υ
+w
∂z
∂x
∂y
∂z
∂t
∂x
∂y
∂z )
(4.35)
242
Chapter 4 Differential Relations for Fluid Flow
This is the differential momentum equation in its full glory, and it is valid for
any fluid in any general motion, particular fluids being characterized by particular viscous stress terms. Note that the last three “convective” terms on the righthand side of each component equation in (4.35) are nonlinear, which complicates
the general mathematical analysis.
Inviscid Flow: Euler’s Equation
Equation (4.35) is not ready to use until we write the viscous stresses in terms
of velocity components. The simplest assumption is frictionless flow τij = 0, for
which Eq. (4.32) reduces to
ρg − ∇p = ρ
dV
dt
(4.36)
This is Euler’s equation for inviscid flow. We show in Sec. 4.9 that Euler’s equation can be integrated along a streamline to yield the frictionless Bernoulli equation, (3.52) or (3.54). The complete analysis of inviscid flow fields, using
continuity and the Bernoulli relation, is given in Chap. 8.
Newtonian Fluid: Navier–Stokes Equations
For a newtonian fluid, as discussed in Sec. 1.7, the viscous stresses are proportional to the element strain rates and the coefficient of viscosity. For incompressible flow, the generalization of Eq. (1.23) to three-dimensional viscous flow is5
τxx = 2μ
∂u
∂x
τyy = 2μ
∂υ
∂y
τzz = 2μ
∂w
∂z
∂u ∂υ
∂w ∂u
τxy = τyx = μ ( + ) τxz = τzx = μ (
+ )
∂y ∂x
∂x
∂z
τyz = τzy = μ (
(4.37)
∂υ ∂w
+
∂z
∂y )
where µ is the viscosity coefficient. Substitution into Eq. (4.35) gives the differential momentum equation for a newtonian fluid with constant density and
viscosity:
ρgx −
ρgy −
ρgz −
∂p
∂ 2u ∂ 2u ∂ 2u
du
+ μ ( 2 + 2 + 2) = ρ
∂x
dt
∂x
∂y
∂z
∂p
∂ 2υ ∂ 2υ ∂ 2υ
dυ
+ μ ( 2 + 2 + 2) = ρ ∂y
dt
∂x
∂y
∂z
(4.38)
∂p
∂ 2w ∂ 2w ∂ 2w
dw
+ μ ( 2 + 2 + 2) = ρ
∂z
dt
∂x
∂y
∂z
These are the incompressible flow Navier–Stokes equations, named after C. L. M.
H. Navier (1785–1836) and Sir George G. Stokes (1819–1903), who are credited
5
When compressibility is significant, additional small terms arise containing the element volume expansion rate and a second coefficient of viscosity; see Refs. 4 and 5 for details.
4.3 The Differential Equation of Linear Momentum 243
with their derivation. They are second-order nonlinear partial differential equations
and are quite formidable, but solutions have been found to a variety of interesting
viscous flow problems, some of which are discussed in Sec. 4.11 and in Chap. 6
(see also Refs. 4 and 5). For compressible flow, see Eq. (2.29) of Ref. 5.
Equations (4.38) have four unknowns: p, u, υ, and w. They should be combined
with the incompressible continuity relation [Eqs. (4.12)] to form four equations
in these four unknowns. We shall discuss this again in Sec. 4.6, which presents
the appropriate boundary conditions for these equations.
Even though the Navier–Stokes equations have only a limited number of
known analytical solutions, they are amenable to fine-gridded computer modeling
[1]. The field of CFD is maturing fast, with many commercial software tools
available. It is possible now to achieve approximate, but realistic, CFD results for
a wide variety of complex two- and three-dimensional viscous flows.
EXAMPLE 4.5
Take the velocity field of Example 4.3, with b = 0 for algebraic convenience
u = a(x2 − y2 )
υ = −2axy w = 0
and determine under what conditions it is a solution to the Navier–Stokes momentum
equations (4.38). Assuming that these conditions are met, determine the resulting pressure ­distribution when z is “up” (gx = 0, gy = 0, gz = –g).
Solution
∙ Assumptions: Constant density and viscosity, steady flow (u and υ independent of
time).
∙ Approach: Substitute the known (u, υ, w) into Eqs. (4.38) and solve for the pressure
gradients. If a unique pressure function p(x, y, z) can then be found, the given solution is exact.
∙ Solution step 1: Substitute (u, υ, w) into Eqs. (4.38) in sequence:
∂p
∂u
∂u
+ μ(2a − 2a + 0) = ρ (u
+ υ ) = 2a2ρ (x3 + xy2 )
∂x
∂x
∂y
∂p
∂υ
∂υ
ρ(0) −
+ μ(0 + 0 + 0) = ρ (u
+ υ ) = 2a2ρ(x2y + y3 )
∂y
∂x
∂y
∂p
∂w
∂w
ρ(−g) −
+ μ(0 + 0 + 0) = ρ (u
+υ
=0
∂z
∂x
∂y )
ρ(0) −
Rearrange and solve for the three pressure gradients:
∂p
= −2a2ρ(x3 + xy2 )
∂x
∂p
= −2a2ρ(x2y + y3 )
∂y
∂p
= −ρg
∂z
(1)
∙ Comment 1: The vertical pressure gradient is hydrostatic. [Could you have predicted
this by noting in Eqs. (4.38) that w = 0?] However, the pressure is velocity-dependent
in the xy plane.
244
Chapter 4 Differential Relations for Fluid Flow
∙ Solution step 2: To determine if the x and y gradients of pressure in Eq. (1) are
­compatible, evaluate the mixed derivative (𝜕2p/𝜕x 𝜕y); that is, cross-differentiate these
two equations:
∂ ∂p
∂
=
[−2a2ρ(x3 + xy2 ) ] = −4a2ρxy
∂y ( ∂x ) ∂y
∂ ∂p
∂
=
[−2a2ρ(x2y + y3 ) ] = −4a2ρxy
∂x ( ∂y ) ∂x
∙ Comment 2: Since these are equal, the given velocity distribution is indeed an exact
solution of the Navier–Stokes equations.
∙ Solution step 3: To find the pressure, integrate Eqs. (1), collect, and compare. Start
with 𝜕p/𝜕x. The procedure requires care! Integrate partially with respect to x, holding
y and z constant:
p=
∫ ∂p∂x dx∣
y,z
= −2a2ρ(x3 + xy2 ) dx∣ y,z = −2a2ρ (
∫
2 2
x4 x y
+
+ f1 (y, z)
4
2 )
(2)
Note that the “constant” of integration f1 is a function of the variables that were not
­integrated. Now differentiate Eq. (2) with respect to y and compare with 𝜕p/𝜕y from Eq. (1):
∂p
∂f1 ∂p
∣ (2) = −2a2ρ x2y +
=
∣ (1) = −2a2ρ(x2y + y3 )
∂y
∂y
∂y
∂f1
∂f1
y4
Compare:
= −2a2ρ y3 or f1 =
dy∣ z = −2a2ρ + f2 (z)
∂y
∂y
4
2 2
4
4
xy
y
x
Collect terms: So far p = −2a2ρ ( +
+ ) + f2 (z) 4
2
4
∫
(3)
This time the “constant” of integration f2 is a function of z only (the variable not integrated). Now differentiate Eq. (3) with respect to z and compare with 𝜕p/𝜕z from Eq. (1):
∂p
df2 ∂p
∣ (3) =
=
∣ (1) = −ρg or
∂z
dz
∂z
f2 = −ρgz + C
(4)
where C is a constant. This completes our three integrations. Combine Eqs. (3) and (4)
to obtain the full expression for the pressure distribution in this flow:
p(x, y, z) = −ρgz − 12 a2ρ(x4 + y4 + 2x2y2 ) + C
Ans. (5)
This is the desired solution. Do you recognize it? Not unless you go back to the beginning and square the velocity components:
u2 + υ2 + w2 = V2 = a2 (x4 + y4 + 2x2y2 )
(6)
Comparing with Eq. (5), we can rewrite the pressure distribution as
p + 12 ρV2 + pgz = C
(7)
∙ Comment: This is Bernoulli’s equation (3.54). That is no accident, because the
velocity distribution given in this problem is one of a family of flows that are solutions to the Navier–Stokes equations and that satisfy Bernoulli’s incompressible equation everywhere in the flow field. They are called irrotational flows, for which curl
V = ∇ × V ≡ 0. This subject is discussed again in Sec. 4.9.
4.4 The Differential Equation of Angular Momentum 245
4.4 The Differential Equation of Angular Momentum
Having now been through the same approach for both mass and linear momentum,
we can go rapidly through a derivation of the differential angular momentum
relation. The appropriate form of the integral angular momentum equation for a
fixed control volume is
∑ Mo =
∂
∂t [
∫
CV
(r × V)ρ d 𝒱 ] +
∫
CS
(r × V)ρ(V · n) dA
(3.59)
We shall confine ourselves to an axis through O that is parallel to the z axis and
passes through the centroid of the elemental control volume. This is shown in
Fig. 4.5. Let θ be the angle of rotation about O of the fluid within the control
volume. The only stresses that have moments about O are the shear stresses τxy
and τyx. We can evaluate the moments about O and the angular momentum terms
about O. A lot of algebra is involved, and we give here only the result:
1 ∂
1 ∂
[ τxy − τyx + 2 ∂x (τxy ) dx − 2 ∂y (τyx ) dy ] dx dy dz
1
d2θ
=
ρ(dx dy dz)(dx2 + dy2 ) 2 12
dt
(4.39)
Assuming that the angular acceleration d2θ/dt2 is not infinite, we can neglect all
higher-order differential terms, which leaves a finite and interesting result:
τxy ≈ τyx
(4.40)
Had we summed moments about axes parallel to y or x, we would have obtained
exactly analogous results:
τxz ≈ τzx
τyz ≈ τzy (4.41)
There is no differential angular momentum equation. Application of the integral
theorem to a differential element gives the result, well known to students of stress
analysis or strength of materials, that the shear stresses are symmetric: τij = τji. This
τ yx + ∂ (τ yx) dy
∂y
τ xy
dy
θ = Rotation
angle
Axis O
Fig. 4.5 Elemental cartesian fixed
control volume showing shear
stresses that may cause a net angular
acceleration about axis O.
dx
τ yx
τ xy + ∂ (τ xy) dx
∂x
246
Chapter 4 Differential Relations for Fluid Flow
is the only result of this section.6 There is no differential equation to remember,
which leaves room in your brain for the next topic, the differential energy equation.
4.5 The Differential Equation of Energy7
We are now so used to this type of derivation that we can race through the energy
equation at a bewildering pace. The appropriate integral relation for the fixed
control volume of Fig. 4.1 is



p
∂
Q − Ws − Wυ = ( eρ d 𝒱) +
e + ρ) ρ(V · n) dA
(3.66)
(
∂t
CV
CS

where Ws = 0 because there can be no infinitesimal shaft protruding into the
control volume. By analogy with Eq. (4.20), the right-hand side becomes, for this
tiny element,
∫
∫


∂
∂
∂
∂
Q − Wυ = [ (ρe) +
(ρuζ ) +
(ρυζ) +
(ρwζ ) ]dx dy dz
∂t
∂x
∂y
∂z
where ζ = e + p/ρ. When we use the continuity equation by analogy with Eq.
(4.21), this becomes


de
Q − Wυ = (ρ
+ V · ∇p + p∇ · V) dx dy dz
dt
(4.42)
Thermal Conductivity; Fourier’s Law
·
To evaluate Q, we neglect radiation and consider only heat conduction through
the sides of the element. Experiments for both fluids and solids show that the
vector heat transfer per unit area, q, is proportional to the vector gradient of
temperature, ∇T. This proportionality is called Fourier’s law of conduction, which
is analogous to Newton’s viscosity law:
or :
q = −k∇T
∂T
∂T
∂T
qx = −k
, qy = −k
, qz = −k
∂x
∂y
∂z
(4.43)
where k is called the thermal conductivity, a fluid property that varies with temperature and pressure in much the same way as viscosity. The minus sign satisfies
the convention that heat flux is positive in the direction of decreasing temperature.
­Fourier’s law is dimensionally consistent, and k has SI units of joules per (sec-­
meter-kelvin) and can be correlated with T in much the same way as Eqs. (1.27)
and (1.28) for gases and liquids, respectively.
Figure 4.6 shows the heat flow passing through the x faces, the y and z heat
flows being omitted for clarity. We can list these six heat flux terms:
6
We are neglecting the possibility of a finite couple being applied to the element by some powerful external force field. See, for example, Ref. 6.
7
This section may be omitted without loss of continuity.
4.5 The Differential Equation of Energy 247
dx
Heat flow per
unit area:
qx = –k ∂T
∂x
qx + ∂ (qx) dx
∂x
dy
wx + ∂ (wx) dx
∂x
wx
Fig. 4.6 Elemental cartesian control
volume showing heat flow and
viscous work rate terms in the x
direction.
Viscous
work rate
per unit
wx = –(u τxx + υ τxy + wτ xz)
area:
Faces
Inlet heat flux
x
qx dy dz
y
qy dx dz
z
qz dx dy
dz
Outlet heat flux
∂
[ qx + ∂x (qx ) dx ] dy dz
∂
[ qy + ∂y (qy ) dy ] dx dz
∂
[ qz + ∂z (qz ) dz ] dx dy
By adding the inlet terms and subtracting the outlet terms, we obtain the net
heat added to the element:

∂
∂
∂
Q = −[ (qx ) +
(qy ) +
(qz ) ]dx dy dz = −∇ · q dx dy dz
(4.44)
∂x
∂y
∂z
As expected, the heat flux is proportional to the element volume. Introducing
Fourier’s law from Eq. (4.43), we have

Q = ∇ · (k∇T ) dx dy dz
(4.45)
The rate of work done by viscous stresses equals the product of the stress
component, its corresponding velocity component, and the area of the element
face. ­Figure 4.6 shows the work rate on the left x face is

Wυ,LF = wx dy dz
where wx = −(uτxx + υ τxy + wτxz )
(4.46)
(where the subscript LF stands for left face) and a slightly different work on the right
face due to the gradient in wx. These work fluxes could be tabulated in exactly the
same manner as the heat fluxes in the previous table, with wx replacing qx, and so on.
After outlet terms are subtracted from inlet terms, the net viscous work rate becomes

∂
∂
Wυ = −[ (uτxx + υ τxy + wτxz ) +
(uτyx + υ τyy + wτyz )
∂x
∂y
∂
+
(uτzx + υτzy + wτzz ) ]dx dy dz
∂z
= −∇ · (V · τij ) dx dy dz
(4.47)
248
Chapter 4 Differential Relations for Fluid Flow
We now substitute Eqs. (4.45) and (4.47) into Eq. (4.43) to obtain one form of
the differential energy equation:
de
+ V · ∇p + p∇ · V = ∇ · (k ∇T ) + ∇ · (V · τij )
dt
where e = û + 12V2 + gz
ρ
(4.48)
A more useful form is obtained if we split up the viscous work term:
∇ · (V · τij ) ≡ V · (∇ · τij ) + Φ
(4.49)
8
where Φ is short for the viscous-dissipation function. For a newtonian incompressible viscous fluid, this function has the form
∂u 2
∂υ 2
∂w 2
∂υ ∂u 2
Φ = μ[ 2 ( ) + 2 ( ) + 2 ( ) + ( + )
∂x
∂y
∂z
∂x ∂y
+(
∂w ∂υ 2
∂u ∂w 2
+ ) +( +
∂y
∂z
∂z
∂x ) ]
(4.50)
Since all terms are quadratic, viscous dissipation is always positive, so a viscous
flow always tends to lose its available energy due to dissipation, in accordance
with the second law of thermodynamics.
Now substitute Eq. (4.49) into Eq. (4.48), using the linear momentum equation
(4.32) to eliminate ∇ . τij. This will cause the kinetic and potential energies to
cancel, leaving a more customary form of the general differential energy equation:
ρ
dû
+ p(∇ · V) = ∇ · (k ∇T ) + Φ dt
(4.51)
This equation is valid for a newtonian fluid under very general conditions of
unsteady, compressible, viscous, heat-conducting flow, except that it neglects
radiation heat transfer and internal sources of heat that might occur during a
chemical or nuclear reaction.
Equation (4.51) is too difficult to analyze except on a digital computer [1]. It
is customary to make the following approximations:
dû ≈ cυ dT cυ , μ, k, ρ ≈ const
(4.52)
Equation (4.51) then takes the simpler form, for ∇ . V = 0,
ρcυ
dT
= k∇ 2T + Φ
dt
(4.53)
which involves temperature T as the sole primary variable plus velocity as a
secondary variable through the total time-derivative operator:
dT ∂T
∂T
∂T
∂T
=
+u
+υ
+w
dt
∂t
∂x
∂y
∂z
8
For further details, see, Ref. 5, p. 72.
(4.54)
4.6 Boundary Conditions for the Basic Equations 249
A great many interesting solutions to Eq. (4.53) are known for various flow
conditions, and extended treatments are given in advanced books on viscous flow
[4, 5] and books on heat transfer [7, 8].
One well-known special case of Eq. (4.53) occurs when the fluid is at rest or has
negligible velocity, where the dissipation Φ and convective terms become negligible:
ρcp
∂T
= k∇ 2T
∂t
(4.55)
The change from cυ to cp is correct and justified by the fact that, when pressure
terms are neglected from a gas flow energy equation [4, 5], what remains is
approximately an enthalpy change, not an internal energy change. This is called
the heat conduction equation in applied mathematics and is valid for solids and
fluids at rest. The solution to Eq. (4.55) for various conditions is a large part of
courses and books on heat transfer.
This completes the derivation of the basic differential equations of fluid motion.
4.6 Boundary Conditions for the Basic Equations
There are three basic differential equations of fluid motion, just derived. Let us
summarize those here:
∂ρ
+ ∇ · (ρV) = 0
∂t
dV
ρ
= ρg − ∇p + ∇ · τij
dt
Continuity:
Momentum:
Energy:
ρ
dû
+ p(∇ · V) = ∇ · (k ∇T ) + Φ dt
(4.56)
(4.57)
(4.58)
where Φ is given by Eq. (4.50). In general, the density is variable, so these three
equations contain five unknowns, ρ, V, p, û, and T. Therefore, we need two additional relations to complete the system of equations. These are provided by data
or algebraic expressions for the state relations of the thermodynamic properties:
ρ = ρ(p, T )
û = û (p, T)
(4.59)
For example, for a perfect gas with constant specific heats, we complete the
system with
ρ=
p
RT
∫
û = cυ dT ≈ cυ T + const
(4.60)
It is shown in advanced books [4, 5] that this system of equations (4.56) to (4.59)
is well posed and can be solved analytically or numerically, subject to the proper
initial and boundary conditions.
First, if the flow is unsteady, there must be an initial condition or initial spatial distribution known for each variable:
At t = 0:
ρ, V, p, û, T = known f (x, y, z)
(4.61)
Thereafter, for all times t to be analyzed, we must know something about the
variables at each boundary enclosing the flow.
250
Chapter 4 Differential Relations for Fluid Flow
Figure 4.7 illustrates the three most common types of boundaries encountered in fluid flow analysis: a solid wall, an inlet or outlet, and a liquid–gas
interface.
First, for a solid, impermeable wall, the most-used boundary condition for
velocity is the no-slip condition in Eq. (4.62). It states that for a fluid in contact
with a solid wall, the velocity of the fluid must equal that of the wall.
Vfluid = Vwall(4.62)
Wall thermal conditions are commonly idealized as being of uniform wall temperature Tw or uniform heat flux qw. If the wall temperature is fixed, the temperature of the fluid can be set to equal that of the wall. If the heat flux at the
wall is constant, Fourier’s law can be applied at the wall to specify the temperature gradient. As a special case, the heat flux equals zero for an adiabatic wall.
They are shown in Eq. (4.63) below
Tfluid = Twall
or
−qw =
∂T
(4.63)
∂n
As indicated in Sec. 1.7, realistic wall slip occurs in rarefied gases when the
Knudsen number is relatively large compared to the characteristic length of the
gaseous flow. The velocity slip and temperature jump are given by
ufluid − uwall ≈ ℓ
Z
∂u
∣wall
∂n
2ζ
k ∂T
Tfluid − Twall ≈ (
ℓ ∣wall )
ζ + 1 μcp ∂n
(4.64)
Liquid–gas interface z = η(x, y, t):
–1
pliq = pgas – Υ(R –1
x + Ry )
dη
wliq = wgas =
dt
Equality of q and τ across interface
Gas
Liquid
Inlet:
known V, p, T
Fig. 4.7 Typical boundary conditions in a viscous heat-conducting
fluid flow analysis.
Outlet:
specified pressure,
∂un /∂n = 0 and
∂T/∂n = 0
Solid contact:
Vfluid = Vwall
Tfluid = Twall or –qw = k ∂T
∂n
Solid impermeable wall
4.6 Boundary Conditions for the Basic Equations 251
where, for the rarefied gas, n is normal to the wall, u is parallel to the wall, ℓ is
the mean free path of the gas [see Eq. (1.37)], and ζ denotes, just this one time,
the specific heat ratio. The above so-called temperature-jump relation for gases
is given here only for completeness and will not be studied (see page 48 of
Ref. 5). A few velocity-jump assignments will be given.
Second, at any outlet section of the flow, the complete distribution of velocity,
pressure, and temperature must be known for all times:
Known V, p, T (4.65)
Outflow boundaries are usually positioned at locations where the flow is approximately unidirectional and where surface stresses are known. For high Reynolds
number flows far from immersed bodies in an external flow or in the fully developed flow out of a pipe, there is no change in any of the velocity components in
the direction across the boundary, and surface forces are Fn = –p and Ft = 0.
This provides the following outflow conditions:
Specified pressure p,
∂un
∂T
= 0 and
= 0
∂n
∂n
(4.66)
Finally, the most complex conditions occur at a liquid–gas interface, or free
­surface, as sketched in Fig. 4.7. Let us denote the interface by
Interface:
z = η(x, y, t)
(4.67)
Then there must be equality of vertical velocity across the interface, so that no
holes appear between liquid and gas:
wliq = wgas =
dη ∂η
∂η
∂η
=
+u
+υ
dt
∂t
∂x
∂y
(4.68)
This is called the kinematic boundary condition.
There must be mechanical equilibrium across the interface. The viscous shear
stresses must balance:
(τzy ) liq = (τzy ) gas
(τzx ) liq = (τzx ) gas
(4.69)
Neglecting the viscous normal stresses, the pressures must balance at the interface
except for surface tension effects:
−1
pliq = pgas − Υ(R−1
x + Ry )
(4.70)
which is equivalent to Eq. (1.33). The radii of curvature can be written in terms
of the free surface position η:
−1
R−1
x + Ry =
∂η/∂x
∂
∂x [ √1 + (∂η/∂x) 2 + (∂η/∂y) 2 ]
+
∂η/∂y
∂
∂y [ √1 + (∂η/∂x) 2 + (∂η/∂y) 2 ]
(4.71)
252
Chapter 4 Differential Relations for Fluid Flow
Finally, the heat transfer must be the same on both sides of the interface, since
no heat can be stored in the infinitesimally thin interface:
(4.72)
(qz ) liq = (qz ) gas
Neglecting radiation, this is equivalent to
∂T
∂T
(k ∂z )liq = (k ∂z )gas
(4.73)
This is as much detail as we wish to give at this level of exposition. Further and
even more complicated details on fluid flow boundary conditions are given in
Refs. 5 and 9.
Simplified Free Surface Conditions
In the introductory analyses given in this book, such as open-channel flows in
Chap. 10, we shall back away from the exact conditions (4.65) to (4.69) and
assume that the upper fluid is an “atmosphere” that merely exerts pressure on the
lower fluid, with shear and heat conduction negligible. We also neglect nonlinear
terms involving the slopes of the free surface. We then have a much simpler and
linear set of conditions at the surface:
∂η
∂ 2η ∂ 2η
pliq ≈ pgas − Υ ( 2 + 2 ) wliq ≈
∂t
∂x
∂y
∂V
∂T
( ∂z )liq ≈ 0 ( z )liq ≈ 0
(4.74)
In many cases, such as open-channel flow, we can also neglect surface tension,
so
pliq ≈ patm
(4.75)
These are the types of approximations that will be used in Chap. 10. The nondimensional forms of these conditions will also be useful in Chap. 5.
Incompressible Flow with Constant Properties
Flow with constant ρ, µ, and k is a basic simplification that will be used, for
example, throughout Chap. 6. The basic equations of motion (4.56) to (4.58)
reduce to
Continuity:
Momentum:
Energy:
∇ · V = 0
ρ
dV
= ρg − ∇p + μ∇2V
dt
ρcp
dT
= k∇2T + Φ
dt
(4.76)
(4.77)
(4.78)
4.6 Boundary Conditions for the Basic Equations 253
Since ρ is constant, there are only three unknowns: p, V, and T. The system is
closed.9 Not only that, the system splits apart: Continuity and momentum are
independent of T. Thus we can solve Eqs. (4.76) and (4.77) entirely separately
for the pressure and velocity, using such boundary conditions as
Solid surface:
V = Vwall
(4.79)
Known V, p
(4.80)
Specified pressure p,
∂un
= 0(4.81)
∂n
Inlet:
Outlet:
Free surface:
p ≈ pa
w≈
∂η
∂t
(4.82)
Later, usually in another course,10 we can solve for the temperature distribution
from Eq. (4.78), which depends on velocity V through the dissipation Φ and the
total time-derivative operator d/dt.
Inviscid Flow Approximations
Chapter 8 assumes inviscid flow throughout, for which the viscosity µ = 0. The
momentum equation (4.77) reduces to
ρ
dV
= ρg − ∇p
dt
(4.83)
This is Euler’s equation; it can be integrated along a streamline to obtain Bernoulli’s equation (see Sec. 4.9). By neglecting viscosity we have lost the secondorder derivative of V in Eq. (4.77); therefore, we must relax one boundary
condition on velocity. The only mathematically sensible condition to drop is the
no-slip condition at the wall. We let the flow slip parallel to the wall but do not
allow it to flow into the wall. The proper inviscid condition is that the normal
velocities must match at any solid surface:
Inviscid flow:
(Vn ) fluid = (Vn ) wall
(4.84)
In most cases the wall is fixed; therefore, the proper inviscid flow condition is
Vn = 0
(4.85)
There is no condition whatever on the tangential velocity component at the wall
in inviscid flow. The tangential velocity will be part of the solution to an inviscid
flow analysis (see Chap. 8).
9
For this system, what are the thermodynamic equivalents to Eq. (4.59)?
Since temperature is entirely uncoupled by this assumption, we may never get around to solving
for it here and may ask you to wait until you take a course on heat transfer.
10
254
Chapter 4 Differential Relations for Fluid Flow
EXAMPLE 4.6
For steady incompressible laminar flow through a long tube, the velocity distribution
is given by
υz = U (1 −
r2
υr = υθ = 0
R2 )
where U is the maximum, or centerline, velocity and R is the tube radius. If the wall
temperature is constant at Tw and the temperature T = T(r) only, find T(r) for this
flow.
Solution
With T = T(r), Eq. (4.78) reduces for steady flow to
ρcpυr
dυz 2
dT k d
dT
=
r
+
μ
( dr ) r dr ( dr )
dr
(1)
But since υr = 0 for this flow, the convective term on the left vanishes. Introduce υz
into Eq. (1) to obtain
dυz 2
4U2μr 2
k d
dT
r
=
−μ
=
−
( dr )
r dr ( dr )
R4
(2)
Multiply through by r/k and integrate once:
2 4
r
μU r
dT
=−
+ C1 dr
kR 4
(3)
Divide through by r and integrate once again:
T=−
μU 2r 4
4kR 4
+ C1 ln r + C2
(4)
Now we are in a position to apply our boundary conditions to evaluate C1 and C2.
First, since the logarithm of zero is –∞, the temperature at r = 0 will be infinite
unless
C1 = 0
(5)
Thus, we eliminate the possibility of a logarithmic singularity. The same thing will
happen if we apply the symmetry condition dT/dr = 0 at r = 0 to Eq. (3). The constant
C2 is then found by the wall-temperature condition at r = R:
T = Tw = −
or
μU2
+ C2
4k
C2 = Tw +
μU2
4k
(6)
4.7 The Stream Function 255
The correct solution is thus
T(r) = Tw +
μU2
r4
1
−
4k (
R4 )
Ans. (7)
which is a fourth-order parabolic distribution with a maximum value T0 = Tw + µU2/
(4k) at the centerline.
4.7 The Stream Function
We have seen in Sec. 4.6 that even if the temperature is uncoupled from our
system of equations of motion, we must solve the continuity and momentum
equations simultaneously for pressure and velocity. The stream function ψ is a
clever device that allows us to satisfy the continuity equation and then solve the
momentum ­equation directly for the single variable ψ. Lines of constant ψ are
streamlines of the flow.
The stream function idea works only if the continuity equation (4.56) can be
reduced to two terms. In general, we have four terms:
Cartesian:
Cylindrical:
∂ρ
∂
∂
∂
+
(ρu) +
(ρυ) +
(ρw) = 0
∂t
∂x
∂y
∂z
(4.86a)
∂ρ 1 ∂
1 ∂
∂
+
(rρυr ) +
(ρυθ ) +
(ρυz ) = 0
r ∂r
r ∂θ
∂t
∂z
(4.86b)
First, let us eliminate unsteady flow, which is a peculiar and unrealistic application of the stream function idea. Reduce either of Eqs. (4.86) to any two terms.
The most common application is incompressible flow in the xy plane:
∂u ∂υ
+
=0
∂x ∂y
(4.87)
This equation is satisfied identically if a function ψ (x, y) is defined such that
Eq. (4.87) becomes
∂ψ
∂ ∂ψ
∂
+
−
≡0
∂x ( ∂y ) ∂y ( ∂x )
(4.88)
Comparison of (4.87) and (4.88) shows that this new function ψ must be defined
such that
u=
or
∂ψ
∂y
V=i
υ=−
∂ψ
∂x
∂ψ
∂ψ
−j
∂y
∂x
(4.89)
256
Chapter 4 Differential Relations for Fluid Flow
Is this legitimate? Yes, it is just a mathematical trick of replacing two variables
(u and υ) by a single higher-order function ψ. The vorticity11 or curl V is an
interesting function:
∂ 2ψ ∂ 2ψ
curl V = −k∇ 2ψ
where
∇ 2ψ = 2 + 2
(4.90)
∂x
∂y
Thus, if we take the curl of the momentum equation (4.77) and utilize Eq. (4.90),
we obtain a single equation for ψ for incompressible flow:
∂ψ ∂
∂ψ ∂
(∇ 2ψ) −
(∇ 2ψ) = ν∇ 2 (∇ 2ψ)
(4.91)
∂y ∂x
∂x ∂y
where ν = µ/ρ is the kinematic viscosity. This is partly a victory and partly a
defeat: Eq. (4.91) is scalar and has only one variable, ψ, but it now contains
fourth-order derivatives and probably will require computer analysis. There will
be four boundary conditions required on ψ. For example, for the flow of a uniform
stream in the x ­direction past a solid body, the four conditions would be
∂ψ
∂ψ
At infinity:
= U∞
=0
∂y
∂x
∂ψ ∂ψ
At the body:
=
=0
(4.92)
∂y
∂x
Many examples of numerical solution of Eqs. (4.91) and (4.92) are given in Ref. 1.
One important application is inviscid, incompressible, irrotational flow12 in
the xy plane, where curl V ≡ 0. Equations (4.90) and (4.91) reduce to
∇ 2ψ =
∂ 2ψ
2
∂x
+
∂ 2ψ
∂y2
=0
(4.93)
This is the second-order Laplace equation (Chap. 8), for which many solutions
and analytical techniques are known. Also, boundary conditions like Eq. (4.92)
reduce to
At infinity:
At the body:
ψ = U∞y + const
ψ = const
(4.94)
It is well within our capability to find some useful solutions to Eqs. (4.93) and
(4.94), which we shall do in Chap. 8.
Geometric Interpretation of ψ
The fancy mathematics above would serve alone to make the stream function
immortal and always useful to engineers. Even better, though, ψ has a beautiful
geometric interpretation: Lines of constant ψ are streamlines of the flow. This
can be shown as follows: From Eq. (1.41) the definition of a streamline in twodimensional flow is
dx dy
=
u
υ
or
u dy − υ dx = 0 streamline
(4.95)
11
12
See Section 4.8.
See Section 4.8.
4.7 The Stream Function 257
Introducing the stream function from Eq. (4.89), we have
∂ψ
∂ψ
dx +
dy = 0 = dψ
∂x
∂y
(4.96)
Thus the change in ψ is zero along a streamline, or
ψ = const along a streamline
(4.97)
Having found a given solution ψ (x, y), we can plot lines of constant ψ to give
the streamlines of the flow.
There is also a physical interpretation that relates ψ to volume flow. From
Fig. 4.8, we can compute the volume flow dQ through an element ds of control
surface of unit depth:
dQ = (V · n) dA = (i
=
∂ψ
∂ψ
dy
dx
−j
· (i
− j ) ds(1)
)
∂y
∂x
ds
ds
∂ψ
∂ψ
dx +
dy = dψ
∂x
∂y
(4.98)
Thus the change in ψ across the element is numerically equal to the volume flow
through the element. The volume flow between any two streamlines in the flow
field is equal to the change in stream function between those streamlines:
Q1→2 =
∫
2
1
(V · n) dA =
2
∫ dψ = ψ
1
2
− ψ1
(4.99)
Further, the direction of the flow can be ascertained by noting whether ψ increases
or decreases. As sketched in Fig. 4.9, the flow is to the right if ψU is greater than
ψL, where the subscripts stand for upper and lower, as before; otherwise the flow
is to the left.
Both the stream function and the velocity potential were invented by the
French mathematician Joseph Louis Lagrange and published in his treatise on
fluid mechanics in 1781.
Control surface
(unit depth
into paper)
dQ = (V • n) d A = dψ
V = iu + jv
dy
ds
Fig. 4.8 Geometric interpretation
of stream function: volume flow
through a differential portion of a
control surface.
dx
n=
dy
dx
i–
j
ds
ds
258
Chapter 4 Differential Relations for Fluid Flow
ψU < ψL
ψU >ψL
Flow
Fig. 4.9 Sign convention for flow in
terms of change in stream function:
(a) flow to the right if ψU is greater;
(b) flow to the left if ψL is greater.
Flow
ψL
ψL
(a)
(b)
EXAMPLE 4.7
If a stream function exists for the velocity field of Example 4.5
u = a(x2 − y2 )
υ = −2axy w = 0
find it, plot it, and interpret it.
Solution
∙ Assumptions: Incompressible, two-dimensional flow.
∙ Approach: Use the definition of stream function derivatives, Eqs. (4.89), to find
ψ (x, y).
∙ Solution step 1: Note that this velocity distribution was also examined in Example
4.3. It satisfies continuity, Eq. (4.87), but let’s check that; otherwise ψ will not exist:
∂u ∂υ
∂
∂
+
=
[a(x2 − y2 ) ] +
(−2axy) = 2ax + (−2ax) ≡ 0 checks
∂x ∂y ∂x
∂y
Thus we are certain that a stream function exists.
∙ Solution step 2: To find ψ, write out Eqs. (4.89) and integrate:
∂ψ
= ax2 − ay2
∂y
∂ψ
υ=−
= −2axy
∂x
u=
(1)
(2)
and work from either one toward the other. Integrate (1) partially
ψ = ax2y −
ay3
+ f(x)
3
(3)
Differentiate (3) with respect to x and compare with (2)
∂ψ
= 2axy + f ′(x) = 2axy
∂x
(4)
Therefore f ′ (x) = 0, or f = constant. The complete stream function is thus found:
ψ = a (x2y −
y3
+ C
3)
Ans. (5)
4.7 The Stream Function 259
To plot this, set C = 0 for convenience and plot the function
3x2y − y3 =
3ψ
a
(6)
for constant values of ψ. The result is shown in Fig. E4.7a to be six 60° wedges of circulating motion, each with identical flow patterns except for the arrows. Once the streamlines are labeled, the flow directions follow from the sign convention of Fig. 4.9. How
can the flow be interpreted? Since there is slip along all streamlines, no streamline can
truly represent a solid surface in a viscous flow. However, the flow could represent the
impingement of three incoming streams at 60, 180, and 300°. This would be a rather unrealistic yet exact solution to the Navier–Stokes equations, as we showed in Example 4.5.
ψ = 2a
a
0
–2a
–a
ψ = 2a
y
ψ = – 2a
E4.7a
–a
0
a
60°
60°
60°
60°
2a
60°
a
x
–a
–2a
The origin is a
stagnation point
Flow around a 60° corner
E4.7b
Flow around a
rounded 60° corner
Incoming stream impinging
against a 120° corner
By allowing the flow to slip as a frictionless approximation, we could let any given
streamline be a body shape. Some examples are shown in Fig. E4.7b.
260
Chapter 4 Differential Relations for Fluid Flow
A stream function also exists in a variety of other physical situations where only
two coordinates are needed to define the flow. Three examples are illustrated here.
Steady Plane Compressible Flow
Suppose now that the density is variable but that w = 0, so that the flow is in
the xy plane. Then the equation of continuity becomes
∂
∂
(ρu) +
(ρυ) = 0
∂x
∂y
(4.100)
We see that this is in exactly the same form as Eq. (4.88). Therefore a compressible flow stream function can be defined such that
ρu =
∂ψ
∂y
ρυ = −
∂ψ
∂x
(4.101)
Again lines of constant ψ are streamlines of the flow, but the change in ψ is now
equal to the mass flow, not the volume flow:

dm = ρ(V · n) dA = dψ
or

m1→2 =
2
∫ ρ(V · n) dA = ψ − ψ 2
1
1
(4.102)
The sign convention on flow direction is the same as in Fig. 4.9. This particular
stream function combines density with velocity and must be substituted into not
only momentum but also the energy and state relations (4.58) and (4.59) with
pressure and temperature as companion variables. Thus the compressible stream
function is not a great victory, and further assumptions must be made to effect
an analytical solution to a typical problem (see, for instance, Ref. 5, Chap. 7).
Incompressible Plane Flow in Polar Coordinates
Suppose that the important coordinates are r and θ, with υz = 0, and that the
density is constant. Then Eq. (4.86b) reduces to
1 ∂
1 ∂
(rυr ) +
(υ ) = 0
r ∂r
r ∂θ θ
(4.103)
After multiplying through by r, we see that this is the analogous form of
Eq. (4.88):
∂ψ
∂ ∂ψ
∂
+
−
=0
(4.104)
∂r ( ∂θ ) ∂θ ( ∂r )
By comparison of (4.103) and (4.104) we deduce the form of the incompressible
polar-coordinate stream function:
υr =
1 ∂ψ
r ∂θ
υθ = −
∂ψ
∂r
(4.105)
Once again lines of constant ψ are streamlines, and the change in ψ is the volume
flow Q1→2 = ψ2 − ψ1. The sign convention is the same as in Fig. 4.9. This type
of stream function is very useful in analyzing flows with cylinders, vortices,
sources, and sinks (Chap. 8).
4.7 The Stream Function 261
Incompressible Axisymmetric Flow
As a final example, suppose that the flow is three-dimensional (υr, υz) but with
no circumferential variations, υθ = 𝜕/𝜕θ = 0 (see Fig. 4.2 for definition of coordinates). Such a flow is termed axisymmetric, and the flow pattern is the same
when viewed on any meridional plane through the axis of revolution z. For incompressible flow, Eq. (4.86b) becomes
1 ∂
∂
(rυr ) +
(υz ) = 0
r ∂r
∂z
(4.106)
This doesn’t seem to work: Can’t we get rid of the one r outside? But when we
­realize that r and z are independent coordinates, Eq. (4.106) can be rewritten as
∂
∂
(rυr ) +
(rυz ) = 0
∂r
∂z
(4.107)
By analogy with Eq. (4.88), this has the form
∂ψ
∂
∂ ∂ψ
−
+
=0
∂r ( ∂z ) ∂z ( ∂r )
(4.108)
By comparing (4.107) and (4.108), we deduce the form of an incompressible
axisymmetric stream function ψ (r, z)
υr = −
1 ∂ψ
r ∂z
υz =
1 ∂ψ
r ∂r
(4.109)
Here again lines of constant ψ are streamlines, but there is a factor (2π) in the
volume flow: Q1→2 = 2π(ψ2 − ψ1 ). The sign convention on flow is the same as
in Fig. 4.9.
EXAMPLE 4.8
Investigate the stream function in polar coordinates
ψ = U sin θ (r −
R2
r)
(1)
where U and R are constants, a velocity and a length, respectively. Plot the streamlines.
What does the flow represent? Is it a realistic solution to the basic equations?
Solution
The streamlines are lines of constant ψ, which has units of square meters per second. Note
that ψ/(UR) is dimensionless. Rewrite Eq. (1) in dimensionless form
ψ
1
= sin θ (η −
η)
UR
η=
r
R
(2)
Of particular interest is the special line ψ = 0. From Eq. (1) or (2) this occurs when (a) θ
= 0 or 180° and (b) r = R. Case (a) is the x axis, and case (b) is a circle of radius R, both
of which are plotted in Fig. E4.8.
262
Chapter 4 Differential Relations for Fluid Flow
For any other nonzero value of ψ it is easiest to pick a value of r and solve for θ:
sin θ =
ψ/(UR)
r/R − R/r
(3)
In general, there will be two solutions for θ because of the symmetry about the y axis. For
example, take ψ/(UR) = +1.0:
Streamlines converge,
high-velocity region
ψ
= +1
UR
r=R
–1
0
0
+1
2
0
0
0
+1
–1
2
–1
Singularity
at origin
E4.8
Guess r/R
3.0
2.5
2.0
1.8
1.7
1.618
Compute θ
22°
158°
28°
152°
42°
138°
53°
127°
64°
116°
90°
This line is plotted in Fig. E4.8 and passes over the circle r = R. Be careful, though, because there is a second curve for ψ/(UR) = +1.0 for small r < R below the x axis:
Guess r/R
0.618
Compute θ
–90°
0.6
0.5
0.4
0.3
0.2
0.1
–70°
–110°
–42°
–138°
–28°
–152°
–19°
–161°
–12°
–168°
–6°
–174°
This second curve plots as a closed curve inside the circle r = R. There is a singularity
of infinite velocity and indeterminate flow direction at the origin. Figure E4.8 shows the
full pattern.
The given stream function, Eq. (1), is an exact and classic solution to the momentum
equation (4.38) for frictionless flow. Outside the circle r = R it represents two-dimensional inviscid flow of a uniform stream past a circular cylinder (Sec. 8.4). Inside the circle it
represents a rather unrealistic trapped circulating motion of what is called a line doublet.
4.8 Vorticity and Irrotationality
The assumption of zero fluid angular velocity, or irrotationality, is a very useful
­simplification. Here we show that angular velocity is associated with the curl of
the local velocity vector.
The differential relations for deformation of a fluid element can be derived by
examining Fig. 4.10. Two fluid lines AB and BC, initially perpendicular at time
4.8 Vorticity and Irrotationality 263
∂u d y d t
∂y
A'
dy +
∂υ dy dt
∂y
Line 2
A
Fig. 4.10 Angular velocity and
strain rate of two fluid lines
deforming in the xy plane.
0
Time: t + dt
C'
dx +
Time t
∂ υ dx dt
∂x
dα
B'
dy
y
dβ
∂u d x d t
∂x
V
B
dx
C
Line 1
x
t, move and deform so that at t + dt they have slightly different lengths A′B′ and
B′C′ and are slightly off the perpendicular by angles dα and dβ. Such deformation
occurs kinematically because A, B, and C have slightly different velocities when
the velocity field V has spatial gradients. All these differential changes in the
motion of A, B, and C are noted in Fig. 4.10.
We define the angular velocity ωz about the z axis as the average rate of
counterclockwise turning of the two lines:
ωz =
1 dα dβ
− )
2 ( dt
dt
(4.110)
But from Fig. 4.10, dα and dβ are each directly related to velocity derivatives in
the limit of small dt:
(∂υ/∂x) dx dt
dα = lim [ tan−1
=
dt→0
dx + (∂u/∂x) dx dt ]
(∂u/∂y) dy dt
dβ = lim [ tan−1
=
dt→0
dy + (∂ υ/∂y) dy dt ]
∂υ
dt
∂x
∂u
dt
∂y
(4.111)
Combining Eqs. (4.110) and (4.111) gives the desired result:
ωz =
1 ∂υ ∂u
−
2 ( ∂x ∂y )
(4.112)
264
Chapter 4 Differential Relations for Fluid Flow
In exactly similar manner we determine the other two rates:
ωx =
1 ∂w ∂υ
− )
2 ( ∂y
∂z
ωy =
1 ∂u ∂w
−
2 ( ∂z
∂x )
(4.113)
The vector ω = iωx + jωy + kωz is thus one-half the curl of the velocity vector
i
1
1 ∂
ω = (curl V) = ∞
2
2 ∂x
u
j
∂
∂y
υ
k
∂
∞ ∂z
w
(4.114)
Since the factor of 12 is annoying, many workers prefer to use a vector twice as
large, called the vorticity:
ζ = 2ω = curl V
(4.115)
Many flows have negligible or zero vorticity and are called irrotational:
curl V ≡ 0
(4.116)
The next section expands on this idea. Such flows can be incompressible or
compressible, steady or unsteady.
We may also note that Fig. 4.10 demonstrates the shear strain rate of the element, which is defined as the rate of closure of the initially perpendicular lines:
dα dβ ∂υ ∂u

εxy =
+
=
+
dt
dt
∂x ∂y
(4.117)
When multiplied by viscosity µ, this equals the shear stress τxy in a newtonian
fluid, as discussed earlier in Eqs. (4.37). Appendix D lists strain rate and vorticity components in cylindrical coordinates.
4.9 Frictionless Irrotational Flows
When a flow is both frictionless and irrotational, pleasant things happen. First,
the momentum equation (4.38) reduces to Euler’s equation:
ρ
dV
= ρg − ∇p
dt
(4.118)
Second, there is a great simplification in the acceleration term. Recall from Sec.
4.1 that acceleration has two terms:
dV ∂V
=
+ (V · ∇)V
dt
∂t
(4.2)
A beautiful vector identity exists for the second term [11]:
(V · ∇)V ≡ ∇( 12V2 ) + ζ × V
where ζ = curl V from Eq. (4.115) is the fluid vorticity.
(4.119)
4.9 Frictionless Irrotational Flows 265
Now combine (4.118) and (4.119), divide by ρ, and rearrange on the left-hand
side. Dot-product the entire equation into an arbitrary vector displacement dr:
∂V
1 2
1
∇p − g · dr = 0
[ ∂t + ∇ ( 2 V ) + ζ × V + ρ
]
(4.120)
Nothing works right unless we can get rid of the third term. We want
(ζ × V) · (dr) ≡ 0
(4.121)
This will be true under various conditions:
1.
2.
3.
4.
V is zero; trivial, no flow (hydrostatics).
ζ is zero; irrotational flow.
dr is perpendicular to ζ × V; this is rather specialized and rare.
dr is parallel to V; we integrate along a streamline (see Sec. 3.5).
Condition 4 is the common assumption. If we integrate along a streamline in frictionless compressible flow and take, for convenience, g = –gk, Eq. (4.120) reduces to
dp
∂V
1
+ g dz = 0
· dr + d ( V 2) +
ρ
∂t
2
(4.122)
Except for the first term, these are exact differentials. Integrate between any two
points 1 and 2 along the streamline:
∫
2
1
∂V
ds +
∂t
∫
2
1
dp 1 2
+ (V 2 − V 12 ) + g(z2 − z1 ) = 0
ρ
2
(4.123)
where ds is the arc length along the streamline. Equation (4.123) is Bernoulli’s
equation for frictionless unsteady flow along a streamline and is identical to Eq.
(3.53). For incompressible steady flow, it reduces to
p 1 2
+ V + gz = constant along streamline
(4.124)
ρ 2
The constant may vary from streamline to streamline unless the flow is also
irrotational (assumption 2). For irrotational flow ζ = 0, the offending term Eq.
(4.121) vanishes regardless of the direction of dr, and Eq. (4.124) then holds all
over the flow field with the same constant.
Velocity Potential
Irrotationality gives rise to a scalar function ϕ similar and complementary to the
stream function ψ. From a theorem in vector analysis [11], a vector with zero curl
must be the gradient of a scalar function
If
∇×V≡0
then
V = ∇ϕ
(4.125)
where ϕ = ϕ (x, y, z, t) is called the velocity potential function. Knowledge of ϕ
thus immediately gives the velocity components
u=
∂ϕ
∂x
υ=
∂ϕ
∂y
w=
∂ϕ
∂z
Lines of constant ϕ are called the potential lines of the flow.
(4.126)
266
Chapter 4 Differential Relations for Fluid Flow
Note that ϕ, unlike the stream function, is fully three-dimensional and is not
limited to two coordinates. It reduces a velocity problem with three unknowns u,
υ, and w to a single unknown potential ϕ; many examples are given in Chap. 8.
The velocity potential also simplifies the unsteady Bernoulli equation (4.122)
because if ϕ exists, we obtain
∂ϕ
∂V
∂
· dr =
(∇ϕ) · dr = d ( )
∂t
∂t
∂t
(4.127)
along any arbitrary direction. Equation (4.122) then becomes a relation between
ϕ and p:
∂ϕ
+
∂t
∫ dpρ + 12 ∣∇ϕ∣
2
+ gz = const
(4.128)
This is the unsteady irrotational Bernoulli equation. It is very important in the
analysis of accelerating flow fields (see Refs. 10 and 15), but the only application
in this text will be in Sec. 9.3 for steady flow.
Orthogonality of Streamlines and Potential Lines
If a flow is both irrotational and described by only two coordinates, ψ and ϕ both
exist, and the streamlines and potential lines are everywhere mutually perpendicular except at a stagnation point. For example, for incompressible flow in the
xy plane, we would have
u=
∂ψ ∂ϕ
=
∂y
∂x
υ=−
∂ψ ∂ϕ
=
∂x
∂y
(4.129)
(4.130)
Can you tell by inspection not only that these relations imply orthogonality but
also that ϕ and ψ satisfy Laplace’s equation?13 A line of constant ϕ would be
such that the change in ϕ is zero:
dϕ =
∂ϕ
∂ϕ
dx +
dy = 0 = u dx + υ dy
∂x
∂y
(4.131)
Solving, we have
dy
u
1
=−
( dx )ϕ=const = − υ
(dy/dx) ψ=const
(4.132)
Equation (4.132) is the mathematical condition that lines of constant ϕ and ψ be
mutually orthogonal. It may not be true at a stagnation point, where both u and
υ are zero, so their ratio in Eq. (4.132) is indeterminate.
13
Equations (4.129) and (4.130) are called the Cauchy–Riemann equations and are studied in
­complex variable theory.
4.9 Frictionless Irrotational Flows 267
Generation of Rotationality14
This is the second time we have discussed Bernoulli’s equation under different
­circumstances (the first was in Sec. 3.5). Such reinforcement is useful, since this
is probably the most widely used equation in fluid mechanics. It requires frictionless flow with no shaft work or heat transfer between sections 1 and 2. The flow
may or may not be irrotational, the former being an easier condition, allowing a
universal Bernoulli constant.
The only remaining question is this: When is a flow irrotational? In other
words, when does a flow have negligible angular velocity? The exact analysis of
fluid rotationality under arbitrary conditions is a topic for advanced study (for
example, Ref. 10, Sec. 8.5; Ref. 9, Sec. 5.2; and Ref. 5, Sec. 2.10). We shall
simply state those results here without proof.
A fluid flow that is initially irrotational may become rotational if
1. There are significant viscous forces induced by jets, wakes, or solid boundaries. In this case Bernoulli’s equation will not be valid in such viscous regions.
2. There are entropy gradients caused by curved shock waves (see Fig. 4.11b).
3. There are density gradients caused by stratification (uneven heating) rather
than by pressure gradients.
4. There are significant noninertial effects such as the earth’s rotation (the
­Coriolis acceleration).
In cases 2 to 4, Bernoulli’s equation still holds along a streamline if friction is
­negligible. We shall not study cases 3 and 4 in this book. Case 2 will be treated
briefly in Chap. 9 on gas dynamics. Primarily we are concerned with case 1,
where rotation is induced by viscous stresses. This occurs near solid surfaces,
where the no-slip condition creates a boundary layer through which the stream
velocity drops to zero, and in jets and wakes, where streams of different velocities
meet in a region of high shear.
Internal flows, such as pipes and ducts, are mostly viscous, and the wall layers
grow to meet in the core of the duct. Bernoulli’s equation does not hold in such
flows unless it is modified for viscous losses.
External flows, such as a body immersed in a stream, are partly viscous and
partly inviscid, the two regions being patched together at the edge of the shear
layer or boundary layer. Two examples are shown in Fig. 4.11. Figure 4.11a shows
a low-speed subsonic flow past a body. The approach stream is irrotational; that
is, the curl of a constant is zero, but viscous stresses create a rotational shear
layer beside and downstream of the body. Generally speaking (see Chap. 7), the
shear layer is laminar, or smooth, near the front of the body and turbulent, or
disorderly, toward the rear. A separated, or deadwater, region usually occurs near
the trailing edge, followed by an unsteady turbulent wake extending far downstream. Some sort of laminar or turbulent viscous theory must be applied to these
viscous regions; they are then patched onto the outer flow, which is frictionless
and irrotational. If the stream Mach number is less than about 0.3, we can combine Eq. (4.126) with the incompressible continuity equation (4.73):
∇ · V = ∇ · (∇ϕ) = 0
14
This section may be omitted without loss of continuity.
268
Chapter 4 Differential Relations for Fluid Flow
Viscous regions where Bernoulli's equation fails:
U
Uniform
approach
flow
(irrotational)
Laminar
boundary
layer
Turbulent
boundary
layer
Separated
flow
Wake
flow
(a)
Curved shock wave introduces rotationality
Viscous regions where Bernoulli is invalid:
Fig. 4.11 Typical flow patterns
illustrating viscous regions patched
onto nearly frictionless regions:
(a) low subsonic flow past a body
(U ≪ a); frictionless, irrotational
potential flow outside the boundary
layer (Bernoulli and Laplace
equations valid); (b) supersonic
flow past a body (U > a);
frictionless, rotational flow outside
the boundary layer (Bernoulli
equation valid, potential flow
invalid).
U
Laminar
boundary
layer
Turbulent
boundary
layer
Slight
separated
flow
Wake
flow
Uniform
supersonic
approach
(irrotational)
(b)
or
∇ 2ϕ = 0 =
∂ 2ϕ
∂x2
+
∂ 2ϕ
∂y2
+
∂ 2ϕ
∂z2
(4.133)
This is Laplace’s equation in three dimensions, there being no restraint on the
number of coordinates in potential flow. A great deal of Chap. 8 will be concerned with solving Eq. (4.133) for practical engineering problems; it holds in
the entire region of Fig. 4.11a outside the shear layer.
Figure 4.11b shows a supersonic flow past a round-nosed body. A curved
shock wave generally forms in front, and the flow downstream is rotational due
to entropy gradients (case 2). We can use Euler’s equation (4.118) in this frictionless region but not potential theory. The shear layers have the same general
character as in Fig. 4.11a except that the separation zone is slight or often absent
and the wake is usually thinner. Theory of separated flow is presently qualitative,
but we can make quantitative estimates of laminar and turbulent boundary layers
and wakes.
4.9 Frictionless Irrotational Flows 269
EXAMPLE 4.9
If a velocity potential exists for the velocity field of Example 4.5
u = a(x2 − y2 )
υ = −2axy w = 0
find it, plot it, and compare with Example 4.7.
Solution
Since w = 0, the curl of V has only one z component, and we must show that it is zero:
(∇ × V) z = 2ωz =
∂υ ∂u
∂
∂
−
=
(−2axy) −
(ax2 − ay2 )
∂x ∂y ∂x
∂y
= −2ay + 2ay = 0
checks
Ans.
The flow is indeed irrotational. A velocity potential exists.
To find ϕ (x, y), set
∂ϕ
u=
= ax2 − ay2
∂x
∂ϕ
υ=
= −2axy
∂y
Integrate (1)
ϕ=
(1)
(2)
ax3
− axy2 + f(y)
3
(3)
Differentiate (3) and compare with (2)
∂ϕ
= −2axy + f ′ (y) = −2axy
∂y
(4)
Therefore f ′ = 0, or f = constant. The velocity potential is
ϕ=
a
–a
ax3
− axy2 + C
3
2a
Ans.
ϕ = –2a
–a
0
y
ϕ = –2a
0
a
2a
x
E4.9
ϕ = 2a a
–2a
0 –a
270
Chapter 4 Differential Relations for Fluid Flow
Letting C = 0, we can plot the ϕ lines in the same fashion as in Example 4.7. The result
is shown in Fig. E4.9 (no arrows on ϕ). For this particular problem, the ϕ lines form the
same pattern as the ψ lines of Example 4.7 (which are shown here as dashed lines) but
are displaced 30°. The ϕ and ψ lines are everywhere perpendicular except at the origin,
a stagnation point, where they are 30° apart. We expected trouble at the stagnation point,
and there is no general rule for determining the behavior of the lines at that point.
4.10 Some Illustrative Incompressible Viscous Flows
Inviscid flows do not satisfy the no-slip condition. They “slip” at the wall but do
not flow through the wall. To look at fully viscous no-slip conditions, we must
attack the complete Navier–Stokes equation (4.77), and the result is usually not
at all irrotational, nor does a velocity potential exist. We look here at three cases:
(1) flow between parallel plates due to a moving upper wall, (2) flow between
parallel plates due to pressure gradient, and (3) flow between concentric cylinders
when the inner one rotates. Other cases will be given as problem assignments or
considered in Chap. 6. Extensive solutions for viscous flows are discussed in Refs.
4 and 5. All flows in this section are viscous and rotational.
Couette Flow between a Fixed and a Moving Plate
Consider two-dimensional incompressible plane (∂/∂z = 0) viscous flow between
parallel plates a distance 2h apart, as shown in Fig. 4.12. We assume that the
plates are very wide and very long, so the flow is essentially axial, u ≠ 0 but υ
= w = 0. The present case is Fig. 4.12a, where the upper plate moves at velocity
V but there is no pressure gradient. Neglect gravity effects. We learn from the
continuity equation (4.76) that
∂u ∂υ ∂w
∂u
+
+
=0=
+0+0
∂x ∂y
∂z
∂x
or
u = u(y) only
Thus there is a single nonzero axial velocity component that varies only across
the channel. The flow is said to be fully developed (far downstream of the
y
u(y)
x
Fig. 4.12 Incompressible viscous
flow between parallel plates: (a) no
pressure gradient, upper plate
moving; (b) pressure gradient ∂p/∂x
with both plates fixed.
Fixed
V
y = +h
u max
u(y)
y = –h
Fixed
Fixed
(a)
(b)
4.10 Some Illustrative Incompressible Viscous Flows 271
entrance). Substitute u = u(y) into the x component of the Navier–Stokes momentum equation (4.77) for two-dimensional (x, y) flow:
ρ (u
or
∂p
∂u
∂u
∂ 2u ∂ 2u
+ υ ) = − + ρgx + μ ( 2 + 2 )
∂x
∂y
∂x
∂x
∂y
2
du
ρ(0 + 0) = 0 + 0 + μ (0 + 2 )
dy
(4.134)
Most of the terms drop out, and the momentum equation reduces to simply
d2u
=0
dy2
or
u = C1y + C2
The two constants are found by applying the no-slip condition at the upper and
lower plates:
At y = +h:
u = V = C1h + C2
At y = −h:
u = 0 = C1 (−h) + C2
or
C1 =
V
2h
and
C2 =
V
2
Therefore the solution for this case (a), flow between plates with a moving upper
wall, is
u=
V
V
y+
2h
2
−h ≤ y ≤ +h
(4.135)
This is Couette flow due to a moving wall: a linear velocity profile with no slip
at each wall, as anticipated and sketched in Fig. 4.12a. Note that the origin has
been placed in the center of the channel for convenience in case (b) which
follows.
What we have just presented is a rigorous derivation of the more informally
­discussed flow of Fig. 1.7 (where y and h were defined differently).
EXAMPLE 4.10
Consider two simple plane flows: (1) A Couette flow shown in Fig. 1.8, the velocity
components are u = ay, v = 0, where a is a constant. (2) A “potential vortex” flow
shown in Fig. 8.3c, which is a useful model for phenomena such as tornadoes and
whirlpools, the velocity components are vr = 0, vθ = K/r, where K is a constant. Investigate the above flows to see if they are rotational.
Solution
We can examine the angular velocity ωz about the vertical axis to determine whether a
flow is rotational or irrotational. For the Couette flow, using Eq. (4.112)
1 ∂υ ∂u
1
ωz = ( − ) = − a
2 ∂x ∂y
2
Ans.
272
Chapter 4 Differential Relations for Fluid Flow
The angular velocity is constant; therefore it is rotational everywhere in the flow.
The “potential vortex” flow is in polar coordinates, from Appendix D
1 1 ∂
1 ∂υr
ωz = (
(rυθ ) −
= 0
r ∂θ )
2 r ∂r
Ans.
It is irrotational.
Comments: The results are somewhat surprising. In the shear-driven plane Couette
flow, the streamlines are straight lines; they look like irrotational. On the other hand,
the streamlines in the “potential vortex” are circulating about the center; they seem rotational. Why does our instinct go so wrong with the fact this time? This is because we
are so used to the rigid body motion in dynamics. When a rigid body is rotating, all the
particles on the rigid body are rotational; when it is in translation, all the particles must be
irrotational. However, for a fluid flow, we need to examine whether or not a small group
of particles on a specific “point” in the flow field is rotating.
Flow Due to Pressure Gradient between Two Fixed Plates
Case (b) is sketched in Fig. 4.12b. Both plates are fixed (V = 0), but the pressure
varies in the x direction. If υ = w = 0, the continuity equation leads to the same
conclusion as case (a)—namely, that u = u(y) only. The x-momentum equation (4.134) changes only because the pressure is variable:
μ
∂p
d2u
=
2
∂x
dy
(4.136)
Also, since υ = w = 0 and gravity is neglected, the y- and z-momentum equations
lead to
∂p
= 0 and
∂y
∂p
= 0 or p = p(x) only
∂z
Thus the pressure gradient in Eq. (4.136) is the total and only gradient:
μ
d2u dp
=
= const < 0
dy2 dx
(4.137)
Why did we add the fact that dp/dx is constant? Recall a useful conclusion from
the theory of separation of variables: If two quantities are equal and one varies
only with y and the other varies only with x, then they must both equal the same
constant. Otherwise they would not be independent of each other.
Why did we state that the constant is negative? Physically, the pressure must
decrease in the flow direction in order to drive the flow against resisting wall
shear stress. Thus the velocity profile u(y) must have negative curvature everywhere, as anticipated and sketched in Fig. 4.12b.
The solution to Eq. (4.137) is accomplished by double integration:
u=
1 dp y2
+ C1y + C2
μ dx 2
4.10 Some Illustrative Incompressible Viscous Flows 273
The constants are found from the no-slip condition at each wall:
At y = ±h:
u=0
or
C1 = 0
and
C2 = −
dp h2
dx 2μ
Thus the solution to case (b), flow in a channel due to pressure gradient, is
u=−
dp h2
y2
1
−
dx 2μ (
h2 )
(4.138)
The flow forms a Poiseuille parabola of constant negative curvature. The maximum velocity occurs at the centerline y = 0:
umax = −
dp h2
dx 2μ
(4.139)
Other (laminar) flow parameters are computed in the following example.
EXAMPLE 4.11
For case (b) in Fig. 4.12b, flow between parallel plates due to the pressure gradient,
compute (a) the wall shear stress, (b) the stream function, (c) the vorticity, (d ) the
velocity potential, and (e) the average velocity.
Solution
All parameters can be computed from the basic solution, Eq. (4.138), by mathematical
­manipulation.
Part (a)
The wall shear follows from the definition of a newtonian fluid, Eq. (4.37):
∂u ∂υ
τw = τxy wall = μ ( + )
∂y ∂x
=±
dp
2μumax
h=∓
dx
h
∣
y = ±h
=μ
dp
y2
∂
h2
−
1
−
∂y [( dx ) ( 2μ ) (
h2 )]
∣
y= ±h
Ans. (a)
The wall shear has the same magnitude at each wall, but by our sign convention of Fig.
4.3, the upper wall has negative shear stress.
Part (b)
Since the flow is plane, steady, and incompressible, a stream function exists:
u=
∂ψ
y2
= umax (1 − 2 )
∂y
h
υ=−
∂ψ
=0
∂x
Integrating and setting ψ = 0 at the centerline for convenience, we obtain
ψ = umax (y −
3h2 )
y3
At the walls, y = ±h and ψ = ±2umaxh/3, respectively.
Ans. (b)
274
Chapter 4 Differential Relations for Fluid Flow
Part (c)
In plane flow, there is only a single nonzero vorticity component:
ζz = (curl V) z =
∂ υ ∂u 2umax
−
= 2 y
∂x
∂y
h
Ans. (c)
The vorticity is highest at the wall and is positive (counterclockwise) in the upper half
and negative (clockwise) in the lower half of the fluid. Viscous flows are typically full of
vorticity and are not at all irrotational.
Part (d)
From part (c), the vorticity is finite. Therefore the flow is not irrotational, and the velocity
potential does not exist.
Ans. (d )
Part (e)
The average velocity is defined as Vav = Q/A, where Q = e u dA over the cross section.
For our particular distribution u(y) from Eq. (4.138), we obtain
Vav =
∫
1
1
u dA =
A
b(2h)
∫
umax (1 −
+h
−h
2)
y2
h
b dy =
2
umax
3
Ans. (e)
In plane Poiseuille flow between parallel plates, the average velocity is two-thirds of the
maximum (or centerline) value. This result could also have been obtained from the stream
function derived in part (b). From Eq. (4.99),
Qchannel = ψupper − ψlower =
2umaxh
2umaxh
4
− (−
= umaxh per unit width
3
3 ) 3
whence Vav = Q/Ab=1 = (4umaxh/3)/(2h) = 2umax/3, the same result.
This example illustrates a statement made earlier: Knowledge of the velocity vector V
[as in Eq. (4.138)] is essentially the solution to a fluid mechanics problem, since all other
flow properties can then be calculated.
Fully Developed Laminar Pipe Flow
Perhaps the most useful exact solution of the Navier–Stokes equation is for incompressible flow in a straight circular pipe of radius R, first studied experimentally
by G. Hagen in 1839 and J. L. Poiseuille in 1840. By fully developed we mean
that the region studied is far enough from the entrance that the flow is purely
axial, υz ≠ 0, while υr and υθ are zero. We neglect gravity and also assume axial
symmetry—that is, ∂/∂θ = 0. The equation of continuity in cylindrical coordinates, Eq. (4.12b), reduces to
∂
(υz ) = 0
or
υz = υz (r)
only
∂z
The flow proceeds straight down the pipe without radial motion. The r-momentum
equation in cylindrical coordinates, Eq. (D.5), simplifies to ∂p/∂r = 0, or p = p(z)
only. The z-momentum equation in cylindrical coordinates, Eq. (D.7), reduces to
ρυz
∂υz
dυz
dp
dp μ d
=− + μ∇ 2υz = − +
r
r dr ( dr )
∂z
dz
dz
4.10 Some Illustrative Incompressible Viscous Flows 275
The convective acceleration term on the left vanishes because of the previously
given continuity equation. Thus the momentum equation may be rearranged as
follows:
dυz
μ d
dp
r
=
= const < 0
r dr ( dr ) dz
(4.140)
This is exactly the situation that occurred for flow between flat plates in Eq.
(4.136). Again the “separation” constant is negative, and pipe flow will look much
like the plate flow in Fig. 4.12b.
Equation (4.140) is linear and may be integrated twice, with the result
υz =
dp r2
+ C1 ln(r) + C2
dz 4μ
where C1 and C2 are constants. The boundary conditions are no slip at the wall
and finite velocity at the centerline:
No slip at r = R: υz = 0 =
dp R2
+ C1 ln(R) + C2
dz 4μ
Finite velocity at r = 0: υz = finite = 0 + C1 ln(0) + C2
To avoid a logarithmic singularity, the centerline condition requires that C1 = 0.
Then, from no slip, C2 = (–dp/dz)(R2/4µ). The final, and famous, solution for
fully developed Hagen–Poiseuille flow is
υz = (−
dp 1
(R2 − r2 ) dz ) 4μ
(4.141)
The velocity profile is a paraboloid with a maximum at the centerline. Just as in
Example 4.11, knowledge of the velocity distribution enables other parameters to
be calculated:
dp R2
Vmax = υz (r = 0) = (− )
dz 4μ
Vavg =
∫
∫
1
1
υz dA = 2
A
πR
Q = υz dA =
∫
R
0
R
0
Vmax (1−
τwall = μ
∣ ∣
∂υz
∂r
max (1− 2 )
r2
R
∫V
r=R
2πr dr =
dp R2
Vmax
= (− )
2
dz 8μ
dp
πR4 Δp
r2
πR4
2
2πr
dr
=
πR
V
=
−
=
avg
8μ ( dz )
8μ L
R2 )
=
4μVavg
R
=
dp
R
R Δp
− )=
(
2
dz
2 L
(4.142)
Note that we have substituted the equality (–dp/dz) = Δp/L, where Δp is the
pressure drop along the entire length L of the pipe.
276
Chapter 4 Differential Relations for Fluid Flow
These formulas are valid as long as the flow is laminar—that is, when the
dimensionless Reynolds number of the flow, ReD = ρVavg(2R)/µ, is less than about
2100. Note also that the formulas do not depend on density, the reason being that
the convective acceleration of this flow is zero.
EXAMPLE 4.12
SAE 10W oil at 20°C flows at 1.1 m3/h through a horizontal pipe with d = 2 cm and
L = 12 m. Find (a) the average velocity, (b) the Reynolds number, (c) the pressure
drop, and (d ) the power required.
Solution
∙ Assumptions: Laminar, steady, Hagen–Poiseuille pipe flow.
∙ Approach: The formulas of Eqs. (4.142) are appropriate for this problem. Note that
R = 0.01 m.
∙ Property values: From Table A.3 for SAE 10W oil, ρ = 870 kg/m3 and µ = 0.104 kg/
(m-s).
∙ Solution steps: The average velocity follows easily from the flow rate and the pipe
area:
Vavg =
Q
πR
2
=
(1.1/3600) m3/s
π(0.01 m) 2
= 0.973
m
s
Ans. (a)
We had to convert Q to m3/s. The (diameter) Reynolds number follows from the average velocity:
Red =
ρVavgd
μ
=
(870 kg/m3 ) (0.973 m/s) (0.02 m)
= 163
0.104 kg/(m-s)
Ans. (b)
This is less than the “transition” value of 2100; so the flow is indeed laminar, and the
­formulas are valid. The pressure drop is computed from the third of Eqs. (4.142):
Q=
4
π(0.01 m) 4 Δp
1.1 m3 πR Δp
=
=
solve for Δp = 97,100 Pa Ans. (c)
3600 s
8μ L
8(0.104 kg/(m-s))(12 m)
When using SI units, the answer returns in pascals; no conversion factors are needed.
­Finally, the power required is the product of flow rate and pressure drop:
1.1
N-m
Power = QΔp = (
m3/s) (97,100 N/m2 ) = 29.7
= 29.7 W Ans. (d )
s
3600
∙ Comments: Pipe flow problems are straightforward algebraic exercises if the data
are compatible. Note again that SI units can be used in the formulas without conversion factors.
Flow between Long Concentric Cylinders
Consider a fluid of constant (ρ, µ) between two concentric cylinders, as in
Fig. 4.13. There is no axial motion or end effect υz = ∂/∂z = 0. Let the inner
cylinder rotate at angular velocity Ωi. Let the outer cylinder be fixed. There is
circular symmetry, so the velocity does not vary with θ and varies only with r.
4.10 Some Illustrative Incompressible Viscous Flows 277
Fixed
ro
Ωi
vθ
ri
Fig. 4.13 Coordinate system for
incompressible viscous flow
between a fixed outer cylinder and
a steadily rotating inner cylinder.
r
Fluid: ρ , μ
The continuity equation for this problem is Eq. (4.12b) with υz = 0:
1 ∂
1 ∂υθ
1 d
(r υr ) +
=0=
(r υr ) or r υr = const
r ∂r
r ∂θ
r dr
Note that υθ does not vary with θ. Since υr = 0 at both the inner and outer cylinders, it follows that υr = 0 everywhere and the motion can only be purely circumferential, υθ = υθ(r). The θ-momentum equation (D.6) becomes
ρ(V · ∇) υθ +
ρυr υθ
υθ
1 ∂p
=−
+ ρgθ + μ ( ∇ 2υθ − 2 )
r
r ∂θ
r
For the conditions of the present problem, all terms are zero except the last.
Therefore, the basic differential equation for flow between rotating cylinders is
∇2υθ =
dυθ
υθ
1 d
r
= 2
(
)
r dr
dr
r
(4.143)
This is a linear second-order ordinary differential equation with the solution
C2
υθ = C1r +
r
The constants are found by the no-slip condition at the inner and outer cylinders:
C2
Outer, at r = ro :
υθ = 0 = C1ro +
ro
C2
Inner, at r = ri :
υθ = Ωiri = C1ri +
ri
The final solution for the velocity distribution is
ro /r − r/ro
Rotating inner cylinder:
υ θ = Ω i ri
(4.144)
ro /ri − ri /ro
The velocity profile closely resembles the sketch in Fig. 4.13. Variations of this
case, such as a rotating outer cylinder, are given in the problem assignments.
278
Chapter 4 Differential Relations for Fluid Flow
Instability of Rotating Inner15 Cylinder Flow
The classic Couette flow solution 16 of Eq. (4.144) describes a physically
satisfying concave, two-dimensional, laminar flow velocity profile as in
Fig. 4.13. The solution is mathematically exact for an incompressible fluid.
However, it becomes unstable at a relatively low rate of rotation of the inner
cylinder, as shown in 1923 in a classic paper by G. I. Taylor [17]. At a
critical value of what is now called the dimensionless Taylor number,
denoted Ta,
Tacrit =
ri (ro − ri ) 3Ω2i
ν2
≈ 1700
(4.145)
the plane flow of Fig. 4.13 vanishes and is replaced by a laminar threedimensional flow pattern consisting of rows of nearly square alternating toroidal vortices. An experimental demonstration of toroidal “Taylor vortices” is
shown in Fig. 4.14 by Dr. Daniel Borrero-Echeverry at Williamette University.
In his experiment, the radius ratio (ri/ro) for the apparatus is 0.91, and it has
an axial aspect ratio (gap to height) of 28. Illustrated in Fig. 4.14 from left
to right, the ratio of the Taylor number to the critical Taylor number (Ta/
Tacrit) for the transition to Taylor rolls is 1.34, 2.01, 8.05, and 29.07, respectively. As the Taylor number increases to 2.01Tacrit and 8.05Tacrit, the vortices
also develop a circumferential periodicity but still laminar. At still higher
Fig. 4.14 Experimental verification
of the instability of flow between a
fixed outer and a rotating inner cylinder. From left to right, the ratio of
the Taylor number to the critical
Taylor number (Ta/Tacrit) is 1.34,
2.01, 8.05, and 29.07, respectively.
(Courtesy of Dr. Daniel BorreroEcheverry at Williamette University)
This instability does not occur if
only the outer cylinder rotates.
15
This section may be omitted without loss of continuity.
Named after M. Couette, whose pioneering paper in 1890 established rotating cylinders as a
method, still used today, for measuring the viscosity of fluids.
16
Problems 279
Ta = 29.07Tacrit, turbulence ensues. This interesting instability reminds us that
the Navier–Stokes equations, being nonlinear, do admit to multiple (nonunique) laminar solutions in addition to the usual instabilities associated with
turbulence and chaotic dynamic systems.
Summary
This chapter complements Chap. 3 by using an infinitesimal control volume to
derive the basic partial differential equations of mass, momentum, and energy for
a fluid. These equations, together with thermodynamic state relations for the fluid
and appropriate boundary conditions, in principle can be solved for the complete
flow field in any given fluid mechanics problem. Except for Chap. 9, in most of
the problems to be studied here an incompressible fluid with constant viscosity
is assumed.
In addition to deriving the basic equations of mass, momentum, and energy,
this chapter introduced some supplementary ideas—the stream function, vorticity,
irrotationality, and the velocity potential—which will be useful in coming chapters, especially Chap. 8. Temperature and density variations will be neglected
except in Chap. 9, where compressibility is studied.
This chapter ended by discussing a few classic solutions for laminar viscous
flows (Couette flow due to moving walls, Poiseuille duct flow due to pressure
gradient, and flow between rotating cylinders). Whole books [4, 5, 9–11, 15]
discuss classic approaches to fluid mechanics, and other texts [6, 12–14]
extend these studies to the realm of continuum mechanics. This does not mean
that all problems can be solved analytically. The new field of computational
fluid dynamics [1] shows great promise of achieving approximate solutions
to a wide variety of flow problems. In addition, when the geometry and boundary conditions are truly complex, experimentation (Chap. 5) is a preferred
alternative.
Problems
Most of the problems herein are fairly straightforward. More
difficult or open-ended assignments are labeled with an asterisk. Problems labeled with a computer icon
may require
the use of a computer. The standard end-of-chapter problems
P4.1 to P4.99 (categorized in the problem list here) are followed by word problems W4.1 to W4.10, fundamentals of
engineering exam problems FE4.1 to FE4.6, and comprehensive problems C4.1 and C4.2.
The acceleration of a fluid
Problem Distribution
P4.1
Section
4.1
4.2
4.3
4.4
4.5
Topic
The acceleration of a fluid
The continuity equation
Linear momentum: Navier–Stokes
Angular momentum: couple stresses
The differential energy equation
Problems
P4.1–P4.8
P4.9–P4.25
P4.26–P4.38
P4.39
P4.40–P4.41
4.6
4.7
4.8 and 4.9
4.7 and 4.9
4.10
4.10
Boundary conditions
Stream function
Velocity potential, vorticity
Stream function and velocity potential
Incompressible viscous flows
Slip flows
P4.42–P4.46
P4.47–P4.55
P4.56–P4.67
P4.68–P4.78
P4.79–P4.96
P4.97–P4.99
An idealized velocity field is given by the formula
V = 4txi − 2t 2yj + 4xzk
Is this flow field steady or unsteady? Is it two- or three-­
dimensional? At the point (x, y, z) = (–1, 1, 0), compute
(a) the acceleration vector and (b) any unit vector normal
to the acceleration.
280
P4.2
Chapter 4 Differential Relations for Fluid Flow
Flow through the converging nozzle in Fig. P4.2
can be approximated by the one-dimensional velocity
­distribution
u ≈ V0 (1 +
2x
υ≈0
L)
P4.6
w≈0
(a) Find a general expression for the fluid acceleration in
the nozzle. (b) For the specific case V0 = 10 ft/s and L =
6 in, compute the acceleration, in g’s, at the entrance and
at the exit.
P4.7
In deriving the continuity equation, we assumed, for simplicity, that the mass flow per unit area on the left face
was just ρu. In fact, ρu varies also with y and z, and thus it
must be different on the four corners of the left face. Account for these variations, average the four corners, and
determine how this might change the inlet mass flow
from ρu dy dz.
Consider a sphere of radius R immersed in a uniform
stream U0, as shown in Fig. P4.7. According to the theory of Chap. 8, the fluid velocity along streamline AB is
given by
R3
V = ui = U0 (1 + 3 ) i
x
V0
P4.3
x=0
y
x
u = U0 (1 + ) v = −U0
L
L
P4.8
U0 y
L
(a) Show that the acceleration vector is purely radial.
(b) For the particular case L = 1.5 m, if the acceleration
at (x, y) = (1 m, 1 m) is 25 m/s2, what is the value of U0?
x
Ut
tanh
2L )
L
Find (a) the fluid acceleration at (x, t) = (L, L/U) and
(b) the time for which the fluid acceleration at x = L is
zero. Why does the fluid acceleration become negative
after ­condition (b)?
w=0
U0 and L are constants
x
When a valve is opened, fluid flows in the expansion
duct of Fig. P4.8 according to the approximation
V = iU (1 −
(a) Sketch a few streamlines in the region 0 < x/L <1 and
0 < y/L < 1, using the method of Sec. 1.11. (b) Find
­expressions for the horizontal and vertical accelerations.
(c) Where is the largest resultant acceleration and its
­numerical value?
P4.5 The velocity field near a stagnation point may be written
in the form
υ=−
R
P4.7
in arbitrary units. At (x, y) = (1, 2), compute (a) the
­accelerations ax and ay, (b) the velocity component in the
direction θ = 40°, (c) the direction of maximum velocity,
and (d ) the direction of maximum acceleration.
A simple flow model for a two-dimensional converging
nozzle is the distribution
U0 x
L
B Sphere
A
x = –4R
A two-dimensional velocity field is given by
u=
y
U0
V = (x2 − y2 + x)i − (2xy + y)j
P4.4
Find (a) the position of maximum fluid acceleration
along AB and (b) the time required for a fluid particle to
travel from A to B.
x=L
x
P4.2
u = 3V0
u (x, t)
P4.8
x=0
x=L
The continuity equation
P4.9
An idealized incompressible flow has the proposed
three-dimensional velocity distribution
V = 4xy2i + f ( y)j − zy2k
Problems 281
Find the appropriate form of the function f( y) that satisfies the continuity relation.
P4.10 A two-dimensional, incompressible flow has the velocity components u = 4y and v = 2x. (a) Find the acceleration components. (b) Is the vector acceleration radial?
(c) Sketch a few streamlines in the first quadrant and
determine if any are straight lines.
P4.11 Derive Eq. (4.12b) for cylindrical coordinates by considering the flux of an incompressible fluid in and out of
the ­elemental control volume in Fig. 4.2.
P4.12 Spherical polar coordinates (r, θ, ϕ) are defined in
Fig. P4.12. The cartesian transformations are
x = r sin θ cos ϕ
y = r sin θ sin ϕ
z = r cos θ
Do not show that the cartesian incompressible continuity
relation [Eq. (4.12a)] can be transformed to the spherical
polar form
1 ∂ 2
1
∂
1
∂
(r υr ) +
(υθ sin θ) +
(υϕ ) = 0
2
∂r
r
sin
θ
∂θ
r
sin
θ
∂ϕ
r
What is the most general form of υr when the flow is
purely radial—that is, υθ and υϕ are zero?
P4.14 For incompressible polar-coordinate flow, what is the
most general form of a purely circulatory motion, υθ =
υθ(r, θ, t) and υr = 0, that satisfies continuity?
P4.15 What is the most general form of a purely radial polar-­
coordinate incompressible flow pattern, υr = υr(r, θ, t)
and υθ = 0, that satisfies continuity?
P4.16 Consider the plane polar-coordinate velocity distribution
vr =
C
r
vθ =
K
r
vz = 0
where C and K are constants. (a) Determine if the equation of continuity is satisfied. (b) By sketching some velocity vector directions, plot a single streamline for C =
K. What might this flow field simulate?
P4.17 An excellent approximation for the two-dimensional
­incompressible laminar boundary layer on the flat surface in Fig. P4.17 is
u ≈ U (2
y
y3 y4
− 2 3 + 4 ) for y ≤ δ
δ
δ
δ
where δ = Cx1/2, C = const
(a) Assuming a no-slip condition at the wall, find an
­expression for the velocity component υ (x, y) for y ≤ δ.
(b) Then find the maximum value of υ at the station x =
1 m, for the particular case of airflow, when U = 3 m/s
and δ = 1.1 cm.
Layer thickness δ (x)
z
υr
y
υϕ
U = constant
U
P
θ
U
r = constant
u(x, y)
u (x, y)
x
0
P4.17
r
υθ
y
ϕ
x
P4.18 A piston compresses gas in a cylinder by moving at constant speed V, as in Fig. P4.18. Let the gas density and
length at t = 0 be ρ0 and L0, respectively. Let the gas velocity vary linearly from u = V at the piston face to u = 0
at x = L. If the gas density varies only with time, find an
­expression for ρ(t).
P4.12
P4.13 For an incompressible plane flow in polar coordinates,
we are given
V = constant
u(x, t)
ρ (t)
υr = r3cos θ + r2sin θ
Find the appropriate form of circumferential velocity for
which continuity is satisfied.
P4.18
x=0
x
x = L(t)
282
Chapter 4 Differential Relations for Fluid Flow
P4.19 A proposed incompressible plane flow in polar coordinates is given by
vr = 2r cos(2θ);
P4.25 An incompressible flow in polar coordinates is given by
υr = K cos θ (1 −
vθ = −2r sin(2θ)
b
r2)
b
(a) Determine if this flow satisfies the equation of contiυθ = −K sin θ (1 + 2 )
nuity. (b) If so, sketch a possible streamline in the first
r
quadrant by finding the velocity vectors at (r, θ) = (1.25,
Does this field satisfy continuity? For consistency, what
20°), (1.0, 45°), and (1.25, 70°). (c) Speculate on what
should the dimensions of constants K and b be? Sketch
this flow might represent.
the surface where υr = 0 and interpret.
P4.20 A two-dimensional incompressible velocity field has u =
–ay
K(1 – e ), for x ≤ L and 0 ≤ y ≤ ∞. What is the most
Linear momentum: Navier–Stokes
general form of υ(x, y) for which continuity is satisfied
*P4.26
Curvilinear, or streamline, coordinates are defined in
and υ = υ0 at y = 0? What are the proper dimensions for
Fig. P4.26, where n is normal to the streamline in the plane
constants K and a?
of the radius of curvature R. Euler’s frictionless momenP4.21 Air flows under steady, approximately one-dimensional
tum equation (4.36) in streamline coordinates becomes
conditions through the conical nozzle in Fig. P4.21. If
the speed of sound is approximately 340 m/s, what is the
∂V
∂V
1 ∂p
+V
=−
+ gs(1)
minimum nozzle-diameter ratio De/D0 for which we can ρ ∂s
∂t
∂s
safely neglect compressibility effects if V0 = (a) 10 m/s
and (b) 30 m/s?
∂θ V2
1 ∂p
−V
−
=−
+ gn(2)
ρ
∂t
R
∂n
Show that the integral of Eq. (1) with respect to s is none
other than our old friend Bernoulli’s equation (3.54).
V0
Ve
n
s, V
z
θ
De
P4.21
D0
P4.22 In an axisymmetric flow, nothing varies with θ, and the
only nonzero velocities are υr and υz (see Fig. 4.2). If the
flow is steady and incompressible and υz = Bz, where B
is constant, find the most general form of υr which satisfies continuity.
P4.23 A tank volume 𝒱 contains gas at conditions (ρ0, p0, T0).
At time t = 0 it is punctured by a small hole of area A.
According to the theory of Chap. 9, the mass flow out of
such a hole is approximately proportional to A and to the
tank pressure. If the tank temperature is assumed constant and the gas is ideal, find an expression for the variation of ­density within the tank.
P4.24 For laminar flow between parallel plates (see Fig. 4.12b),
the flow is two-dimensional (υ ≠ 0) if the walls are
­porous. A special case solution is u = (A − Bx) (h2 − y2 ),
where A and B are constants. (a) Find a general formula
for velocity υ if υ = 0 at y = 0. (b) What is the value of
the constant B if υ = υw at y = +h?
y
Streamline
R
x
P4.26
P4.27 A frictionless, incompressible steady flow field is given
by
V = 2xyi − y2j
in arbitrary units. Let the density be ρ0 = constant and
neglect gravity. Find an expression for the pressure
­
­gradient in the x direction.
P4.28 For the velocity distribution of Prob. 4.10, (a) check continuity. (b) Are the Navier–Stokes equations valid? (c) If
so, determine p(x, y) if the pressure at the origin is p0.
P4.29 Consider a steady, two-dimensional, incompressible
flow of a newtonian fluid in which the velocity field is
known: u = –2xy, υ = y2 – x2, w = 0. (a) Does this
flow satisfy conservation of mass? (b) Find the pressure
field, p(x, y) if the pressure at the point (x = 0, y = 0) is
equal to pa.
Problems 283
P4.30 For the velocity distribution of Prob. P4.4, determine if
(a) the equation of continuity and (b) the Navier–Stokes
equation are satisfied. (c) If the latter is true, find the
pressure distribution p(x, y) when the pressure at the origin equals po.
P4.31 According to potential theory (Chap. 8) for the flow
­approaching a rounded two-dimensional body, as in
Fig. P4.31, the velocity approaching the stagnation point
is given by u = U(1 – a2/x2), where a is the nose radius
and U is the velocity far upstream. Compute the value
and position of the maximum viscous normal stress
along this streamline.
Stagnation
point
(u = 0)
P4.34 A proposed three-dimensional incompressible flow field
has the following vector form:
V = Kxi + Kyj − 2Kzk
(a) Determine if this field is a valid solution to continuity
and Navier–Stokes. (b) If g = –gk, find the pressure field
p(x, y, z). (c) Is the flow irrotational?
P4.35 From the Navier–Stokes equations for incompressible
flow in polar coordinates (App. D for cylindrical coordinates), find the most general case of purely circulating
motion υθ(r), υr = υz = 0, for flow with no slip between
two fixed concentric cylinders, as in Fig. P4.35.
y
a
υθ (r)
x
0
r
r=a
No slip
P4.31
r=b
P4.35
Is this also the position of maximum fluid deceleration?
Evaluate the maximum viscous normal stress if the fluid
is SAE 30 oil at 20°C, with U = 2 m/s and a = 6 cm.
P4.32 The answer to Prob. P4.14 is υθ = f(r) only. Do not reveal
this to your friends if they are still working on Prob.
P4.14. Show that this flow field is an exact solution to
the Navier–Stokes equations (4.38) for only two special
cases of the function f(r). Neglect gravity. Interpret these
two cases physically.
P4.33 Consider incompressible flow at a volume rate Q toward
a drain at the vertex of a 45° wedge of width b, as in
Fig. P4.33. Neglect gravity and friction and assume purely
radial inflow. (a) Find an expression for υr(r). (b) Show
that the viscous term in the r-momentum equation is zero.
(c) Find the pressure distribution p(r) if p = po at r = R.
P4.36 A constant-thickness film of viscous liquid flows in
­laminar motion down a plate inclined at angle θ, as in
Fig. P4.36. The velocity profile is
u = Cy(2h − y)
(a) Find the constant C in terms of the specific weight
and viscosity and the angle θ. (b) Find the volume flow
rate Q per unit width in terms of these parameters.
y
g
h
θ =π /4
u(y)
Q
r
θ
P4.33
Drain
υ=w=0
θ
P4.36
x
*P4.37 A viscous liquid of constant ρ and µ falls due to gravity
between two plates a distance 2h apart, as in Fig. P4.37.
The flow is fully developed, with a single velocity
284
Chapter 4 Differential Relations for Fluid Flow
­component w = w(x). There are no applied pressure gradients, only gravity. Solve the Navier–Stokes equation
for the ­velocity profile between the plates.
Tw
y=h
y
h
h
y=0
x
P4.41
z, w
u(y)
T(y)
x
Tw
Boundary conditions
P4.39 Reconsider the angular momentum balance of Fig. 4.5
by adding a concentrated body couple Cz about the z axis
[6]. Determine a relation between the body couple and
shear stress for equilibrium. What are the proper dimensions for Cz? (Body couples are important in continuous
media with microstructure, such as granular materials.)
P4.42 Suppose we wish to analyze the rotating, partly full cylinder of Fig. 2.23 as a spin-up problem, starting from
rest and continuing until solid-body rotation is achieved.
What are the appropriate boundary and initial conditions
for this problem?
P4.43 For the draining liquid film of Fig. P4.36, what are the
­appropriate boundary conditions (a) at the bottom y = 0
and (b) at the surface y = h?
P4.44 Suppose that we wish to analyze the sudden pipe expansion flow of Fig. P3.59, using the full continuity and
Navier–Stokes equations. What are the proper boundary
conditions to handle this problem?
P4.45 For the sluice gate problem of Example 3.10, list all the
boundary conditions needed to solve this flow exactly
by, say, computational fluid dynamics.
P4.46 Fluid from a large reservoir at temperature T0 flows into
a circular pipe of radius R. The pipe walls are wound
with an electric resistance coil that delivers heat to the
fluid at a rate qw (energy per unit wall area). If we wish
to analyze this problem by using the full continuity, Navier–Stokes, and energy equations, what are the proper
boundary conditions for the analysis?
The differential energy equation
Stream function
P4.37
P4.38 Show that the incompressible flow distribution, in cylindrical coordinates,
vr = 0
vθ = Crn
vz = 0
where C is a constant, (a) satisfies the Navier–Stokes
equation for only two values of n. Neglect gravity. (b)
Knowing that p = p(r) only, find the pressure distribution
for each case, assuming that the pressure at r = R is p0.
What might these two cases represent?
Angular momentum: couple stresses
P4.40 For pressure-driven laminar flow between parallel plates
(see Fig. 4.12b), the velocity components are u = U(1–
y2/h2), υ = 0, and w = 0, where U is the centerline
­velocity. In the spirit of Ex. 4.6, find the temperature
distribution T(y) for a constant wall temperature Tw.
P4.41 As mentioned in Sec. 4.10, the velocity profile for laminar flow between two plates, as in Fig. P4.41, is
u=
4umax y(h − y)
h2
υ=w=0
If the wall temperature is Tw at both walls, use the
­incompressible flow energy equation (4.75) to solve for
the temperature distribution T(y) between the walls for
steady flow.
P4.47 A two-dimensional incompressible flow is given by the
velocity field V = 3yi + 2xj, in arbitrary units. Does
this flow satisfy continuity? If so, find the stream function ψ(x, y) and plot a few streamlines, with arrows.
P4.48 Consider the following two-dimensional incompressible
flow, which clearly satisfies continuity:
u = U0 = constant, υ = V0 = constant
Find the stream function ψ(r, θ) of this flow using polar
coordinates.
P4.49 Investigate the stream function ψ = K(x2 – y2), K =
­constant. Plot the streamlines in the full xy plane, find
any stagnation points, and interpret what the flow could
­represent.
Problems 285
P4.50 In 1851, George Stokes (of Navier–Stokes fame) solved
the problem of steady incompressible low-Reynoldsnumber flow past a sphere, using spherical polar coordinates (r, θ) [Ref. 5, page 168]. In these coordinates, the
equation of continuity is
∂ 2
∂
(r υr sin θ) +
(r υθ sin θ ) = 0
∂r
∂θ
(a) Does a stream function exist for these coordinates?
(b) If so, find its form.
P4.51 The velocity profile for pressure-driven laminar flow
­between parallel plates (see Fig. 4.12b) has the form u =
C(h2 – y2), where C is a constant. (a) Determine if a
stream function exists. (b) If so, find a formula for the
stream ­function.
P4.52 A two-dimensional, incompressible, frictionless fluid is
guided by wedge-shaped walls into a small slot at the
­origin, as in Fig. P4.52. The width into the paper is b,
θ = π /4
P4.55 The proposed flow in Prob. P4.19 does indeed satisfy the
equation of continuity. Determine the polar-coordinate
stream function of this flow.
Velocity potential, vorticity
P4.56 Investigate the velocity potential ϕ = Kxy, K = constant.
Sketch the potential lines in the full xy plane, find any
stagnation points, and sketch in by eye the orthogonal
streamlines. What could the flow represent?
P4.57 A two-dimensional incompressible flow field is defined
by the velocity components
x y
u = 2V ( − )
L L
r
Slot
θ=0
P4.52
and the volume flow rate is Q. At any given distance r
from the slot, the flow is radial inward, with constant velocity. Find an expression for the polar-coordinate stream
function of this flow.
P4.53 For the fully developed laminar pipe flow solution of
Eq. (4.137), find the axisymmetric stream function ψ(r,
z). Use this result to determine the average velocity V =
Q/A in the pipe as a ratio of umax.
P4.54 An incompressible stream function is defined by
∂ϕ
∂r
υθ =
1 ∂ϕ
r ∂θ
Finally show that ϕ as defined here satisfies Laplace’s
equation in polar coordinates for incompressible flow.
P4.59 Consider the two-dimensional incompressible velocity
potential ϕ = xy + x2 – y2. (a) Is it true that ∇2ϕ = 0, and,
if so, what does this mean? (b) If it exists, find the stream
function ψ(x, y) of this flow. (c) Find the equation of the
streamline that passes through (x, y) = (2, 1).
P4.60 Liquid drains from a small hole in a tank, as shown in
Fig. P4.60, such that the velocity field set up is given by
υr ≈ 0, υz ≈ 0, υθ = KR2/r, where z = H is the depth of the
water far from the hole. Is this flow pattern rotational or
irrotational? Find the depth zC of the water at the radius
r = R.
z
patm
U
ψ (x, y) = 2 (3x2y − y3 )
L
where U and L are (positive) constants. Where in this
­chapter are the streamlines of this flow plotted? Use
this stream function to find the volume flow Q passing
through the rectangular surface whose corners are defined by (x, y, z) = (2L, 0, 0), (2L, 0, b), (0, L, b), and (0,
L, 0). Show the direction of Q.
y
L
where V and L are constants. If they exist, find the stream
function and velocity potential.
P4.58 Show that the incompressible velocity potential in plane
polar coordinates ϕ(r, θ) is such that
υr =
vr
υ = −2V
r
z=H
zC?
z=0
r=R
P4.60
286
Chapter 4 Differential Relations for Fluid Flow
P4.61 An incompressible stream function is given by
ψ = aθ + br sin θ. (a) Does this flow have a velocity
­potential? (b) If so, find it.
P4.62 Show that the linear Couette flow between plates in Fig.
1.8 has a stream function but no velocity potential. Why
is this so?
P4.63 Find the two-dimensional velocity potential ϕ(r, θ) for
the polar-coordinate flow pattern υr = Q/r, υθ = K/r,
where Q and K are constants.
P4.64 Show that the velocity potential ϕ(r, z) in axisymmetric
cylindrical coordinates (see Fig. 4.2) is defined such that
υr =
∂ϕ
∂r
υz =
K = const
Find the stream function for this flow, sketch some streamlines and potential lines, and interpret the flow pattern.
P4.67 A stream function for a plane, irrotational, polar-coordinate flow is
ψ = Cθ − K ln r
y = 1.1 m
ψ = 1.9552 m2/s
C and K = const
Find the velocity potential for this flow. Sketch some
streamlines and potential lines, and interpret the flow
­pattern.
Stream function and velocity potential
P4.68 For the velocity distribution of Prob. P4.4, (a) determine
if a velocity potential exists, and (b), if it does, find an
­expression for ϕ(x, y) and sketch the potential line which
passes through the point (x, y) = (L/2, L/2).
P4.69 A steady, two-dimensional flow has the following polarcoordinate velocity potential:
α?
1.7308 m2/s
y = 1.0 m
1.7978
x = 1.5 m
x = 1.6 m
P4.70
P4.71 Consider the following two-dimensional function f(x, y):
f = Ax3 + Bxy2 + Cx2 + D
where A > 0
(a) Under what conditions, if any, on (A, B, C, D) can
this function be a steady plane-flow velocity potential?
(b) If you find a ϕ(x, y) to satisfy part (a), also find the
associated stream function ψ(x, y), if any, for this flow.
P4.72 Water flows through a two-dimensional narrowing wedge
at 9.96 gal/min per meter of width into the paper
(Fig. P4.72). If this inward flow is purely radial, find an
expression, in SI units, for (a) the stream function and
(b) the velocity potential of the flow. Assume one-­
dimensional flow. The included angle of the wedge is 45°.
Drain
ϕ = C r cos θ + K ln r
where C and K are constants. Determine the stream function ψ(r, θ ) for this flow. For extra credit, let C be a velocity scale U and let K = UL, sketch what the flow might
represent.
P4.70 A CFD model of steady two-dimensional incompressible
flow has printed out the values of stream function ψ(x, y),
in m2/s, at each of the four corners of a small 10-cm-by10-cm cell, as shown in Fig. P4.70. Use these numbers to
2.0206
V?
∂ϕ
∂z
Further show that for incompressible flow this potential
satisfies Laplace’s equation in (r, z) coordinates.
P4.65 Consider the function f = ay – by3. (a) Could this represent a realistic velocity potential? Extra credit: (b) Could
it represent a stream function?
P4.66 A plane polar-coordinate velocity potential is defined by
K cos θ
ϕ=
r
estimate the resultant velocity in the center of the cell
and its angle α with respect to the x axis.
Q
r
P4.72
P4.73 A CFD model of steady two-dimensional incompressible flow has printed out the values of velocity potential
Problems 287
ϕ(x, y), in m2/s, at each of the four corners of a small
10-cm-by-10-cm cell, as shown in Fig. P4.73. Use these
numbers to estimate the resultant velocity in the center
of the cell and its angle α with respect to the x axis.
y = 1.1 m
ϕ = 4.8338 m2/s
5.0610
V?
in the upper half plane. (b) Prove that a stream function
exists, and then find ψ(x, y), using the hint that ∫dx/(a2 +
x2) = (1/a)tan–1(x/a).
P4.77 Outside an inner, intense-activity circle of radius R, a
tropical storm can be simulated by a polar-coordinate
velocity potential ϕ(r, θ) = UoR θ, where Uo is the wind
velocity at radius R. (a) Determine the velocity components outside r = R. (b) If, at R = 25 mi, the velocity is
100 mi/h and the pressure 99 kPa, calculate the velocity
and pressure at r = 100 mi.
P4.78 An incompressible, irrotational, two-dimensional flow
has the following stream function in polar coordinates:
ψ = A rn sin (nθ)
α?
where A and n are constants.
Find an expression for the velocity potential of this flow.
Incompressible viscous flows
*P4.79 Study the combined effect of the two viscous flows in
Fig. 4.12. That is, find u(y) when the upper plate moves
at speed V and there is also a constant pressure gradient
y = 1.0 m
(dp/dx). Is superposition possible? If so, explain why.
x = 1.5 m
x = 1.6 m
Plot representative velocity profiles for (a) zero, (b) positive, and (c) negative pressure gradients for the same
P4.73
upper-wall speed V.
*P4.80
Oil, of density ρ and viscosity µ, drains steadily down
P4.74 Consider the two-dimensional incompressible polar-­
the
side of a vertical plate, as in Fig. P4.80. After a decoordinate velocity potential
velopment region near the top of the plate, the oil film
ϕ = Br cos θ + B L θ
will ­become independent of z and of constant thickness
δ. ­Assume that w = w(x) only and that the atmosphere
where B is a constant and L is a constant length scale.
offers no shear resistance to the surface of the film. (a)
(a) What are the dimensions of B? (b) Locate the only
Solve the Navier–Stokes equation for w(x), and sketch
­stagnation point in this flow field. (c) Prove that a stream
its approximate shape. (b) Suppose that film thickness δ
function exists and then find the function ψ(r, θ).
and the slope of the velocity profile at the wall [∂w/∂x]wall
P4.75 Given the following steady axisymmetric stream function:
are measured with a laser-Doppler anemometer (Chap.
6). Find an e­ xpression for oil viscosity µ as a function of
B
r4
ψ = (r2 − 2 ) where B and R are constants
(ρ, δ, g, [∂w/∂x]wall).
2
2R
4.9038 m2/s
5.1236
Plate
valid in the region 0 ⩽ r ⩽ R and 0 ⩽ z ⩽ L. (a) What are
the dimensions of the constant B? (b) Show whether this
flow possesses a velocity potential, and, if so, find it.
(c) What might this flow represent? Hint: Examine the
­axial velocity vz.
*P4.76 A two-dimensional incompressible flow has the velocity
potential
Oil film
Air
δ
g
ϕ = K(x2 − y2 ) + C ln(x2 + y2 )
where K and C are constants. In this discussion, avoid the
origin, which is a singularity (infinite velocity). (a) Find
the sole stagnation point of this flow, which is somewhere
z
P4.80
x
288
Chapter 4 Differential Relations for Fluid Flow
P4.81 Modify the analysis of Fig. 4.13 to find the velocity uθ *P4.84 Consider a viscous film of liquid draining uniformly
when the inner cylinder is fixed and the outer cylinder
down the side of a vertical rod of radius a, as in Fig.
P4.84. At some distance down the rod the film will ap­rotates at angular velocity Ω0. May this solution be added
proach a terminal or fully developed draining flow of
to Eq. (4.140) to represent the flow caused when both inner and outer cylinders rotate? Explain your conclusion.
constant outer ­radius b, with υz = υz(r), υθ = υr = 0. As*P4.82 A solid circular cylinder of radius R rotates at angular vesume that the atmosphere offers no shear resistance to
the film motion. Derive a differential equation for υz,
locity Ω in a viscous incompressible fluid that is at rest far
from the cylinder, as in Fig. P4.82. Make simplifying asstate the proper boundary conditions, and solve for the
sumptions and derive the governing differential equation
film velocity distribution. How does the film radius b
and boundary conditions for the velocity field υθ in the
relate to the total film volume flow rate Q?
fluid. Do not solve unless you are obsessed with this problem. What is the steady-state flow field for this problem?
υθ (r, θ , t)
r
Q
r
z
θ
Ω
Fully
developed
region
r=R
P4.82
P4.83 The flow pattern in bearing lubrication can be illustrated
by Fig. P4.83, where a viscous oil (ρ, µ) is forced into the
gap h(x) between a fixed slipper block and a wall moving
at velocity U. If the gap is thin, h ≪ L, it can be shown
that the pressure and velocity distributions are of the
form p = p(x), u = u(y), υ = w = 0. Neglecting gravity,
reduce the ­Navier–Stokes equations (4.38) to a single
differential equation for u(y). What are the proper boundary ­conditions? Integrate and show that
u=
pa
µa ≈ 0
y
1 dp 2
(y − yh) + U (1 − )
2μ dx
h
where h = h(x) may be an arbitrary, slowly varying gap
width. (For further information on lubrication theory,
see Ref. 16.)
a
b
Film
µ
ρ
υz
P4.84
P4.85 A flat plate of essentially infinite width and breadth oscillates sinusoidally in its own plane beneath a viscous
fluid, as in Fig. P4.85. The fluid is at rest far above the
plate. Making as many simplifying assumptions as you
can, set up the governing differential equation and
boundary conditions for finding the velocity field u in
the fluid. Do not solve (if you can solve it immediately,
you might be able to get exempted from the balance of
this course with credit).
y
Incompressible
viscous
fluid
Oil
inlet
Fixed slipper
block
h0
x
h (x)
Moving wall
P4.83
y
Oil
outlet
u (y)
u (x, y, z, t)?
x
Plate velocity:
h1
U
P4.85
U0 sin ωt
P4.86 SAE 10 oil at 20°C flows between parallel plates 8 mm
apart, as in Fig. P4.86. A mercury manometer, with wall
Problems 289
pressure taps 1 m apart, registers a 6-cm height, as
shown. Estimate the flow rate of oil per unit width for
this condition.
P4.89 Oil flows steadily between two fixed plates that are 2
inches apart. When the pressure gradient is 3200 pascals
per ­meter, the average velocity is 0.8 m/s. (a) What is the
flow rate per meter of width? (b) What oil in Table A.4
fits this data? (c) Can we be sure that the flow is laminar?
SAE 10
8 mm
Q
P4.90 It is desired to pump ethanol at 20°C through 25 m of
oil
straight smooth tubing under laminar-flow conditions,
Red = ρVd/µ < 2300. The available pressure drop is
10 kPa. (a) What is the maximum possible mass flow, in
6 cm
kg/h? (b) What is the appropriate diameter?
Mercury
*P4.91 Analyze fully developed laminar pipe flow for a powerlaw fluid, τ = C(dvz /dr)n, for n ≠ 1, as in Prob. P1.46.
1
m
P4.86
(a) ­Derive an expression for vz(r). (b) For extra credit,
plot the velocity profile shapes for n = 0.5, 1, and 2.
P4.87 SAE 30W oil at 20°C flows through the 9-cm-diameter
[Hint: In Eq. (4.136), replace µ(dvz/dr) with τ.]
pipe in Fig. P4.87 at an average velocity of 4.3 m/s.
P4.92 A tank of area A0 is draining in laminar flow through a
pipe of diameter D and length L, as shown in Fig. P4.92.
D = 9 cm
­Neglecting the exit jet kinetic energy and assuming the
pipe flow is driven by the hydrostatic pressure at its
V
SAE 30W oil
­entrance, derive a formula for the tank level h(t) if its
initial level is h0.
h
Hg
Area Ao
2.5 m
P4.87
(a) Verify that the flow is laminar. (b) Determine the
volume flow rate in m3/h. (c) Calculate the expected
reading h of the mercury manometer, in cm.
P4.88 The viscous oil in Fig. P4.88 is set into steady motion by
a concentric inner cylinder moving axially at velocity U
­inside a fixed outer cylinder. Assuming constant pressure and density and a purely axial fluid motion, solve
Eqs. (4.38) for the fluid velocity distribution υz(r). What
are the proper boundary conditions?
Fixed outer cylinder
r
b
vz(r)
a
Oil: ρ, µ
P4.88
vz
U
h(t)
ρ, µ
D, L
V(t)
P4.92
P4.93 A number of straight 25-cm-long microtubes of diameter
d are bundled together into a “honeycomb” whose total
cross-sectional area is 0.0006 m2. The pressure drop
from entrance to exit is 1.5 kPa. It is desired that the total
volume flow rate be 5 m3/h of water at 20°C. (a) What is
the ­appropriate microtube diameter? (b) How many microtubes are in the bundle? (c) What is the Reynolds
number of each microtube?
P4.94 A long, solid cylinder rotates steadily in a very viscous
fluid, as in Fig. P4.94. Assuming laminar flow, solve the
Navier–Stokes equation in polar coordinates to determine the resulting velocity distribution. The fluid is at
rest far from the cylinder. [Hint: The cylinder does not
induce any radial motion.]
Chapter 4 Differential Relations for Fluid Flow
290
P4.96 Use the data of Prob. P1.40, with the inner cylinder rotating
and outer cylinder fixed, and calculate (a) the inner shear
stress. (b) Determine whether this flow pattern is stable.
[Hint: The shear stress in (r, θ) coordinates is not like plane
flow.
r
ρ, µ
Slip flows
R
Ω
P4.94
*P4.95 Two immiscible liquids of equal thickness h are being
sheared between a fixed and a moving plate, as in
Fig. P4.95. Gravity is neglected, and there is no variation
with x. Find an expression for (a) the velocity at the interface and (b) the shear stress in each fluid. Assume steady
laminar flow.
V
y
h
P4.97 For Couette flow between a moving and a fixed plate,
Fig. 4.12a, solve continuity and Navier–Stokes to find
the velocity distribution when there is slip at both walls.
P4.98 For the pressure-gradient flow between two parallel
plates of Fig. 4.12b, reanalyze for the case of slip flow at
both walls. Use the simple slip condition uwall = ℓ(du/
dy)wall, where ℓ is the mean free path of the fluid. (a)
Sketch the expected ­velocity profile. (b) Find an expression for the velocity distribution. (c) Find the volume
flow rate ­between the plates.
P4.99 For the pressure-gradient flow in a circular tube in
Sec. 4.10, reanalyze for the case of slip flow at the
wall. Use the simple slip condition υz,wall = ℓ(dvz/
dr)wall, where ℓ is the mean free path of the fluid. (a)
Sketch the ­expected velocity profile. (b) Find an expression for the velocity distribution. (c) Find the volume flow rate through the tube. (d) Is the volume flow
rate increased or decreased, compared with that calculated using the no-slip velocity at the wall?
ρ 2, µ 2
h
ρ 1, µ 1
x
Fixed
P4.95
Word Problems
W4.1 The total acceleration of a fluid particle is given by
Eq. (4.2) in the Eulerian[?] system, where V is a known
function of space and time. Explain how we might evaluate particle acceleration in the Lagrangian[?] frame,
where particle position r is a known function of time
and initial position, r = fcn(r0, t). Can you give an illustrative example?
W4.2 Is it true that the continuity relation, Eq. (4.6), is valid for
both viscous and inviscid, newtonian and nonnewtonian,
compressible and incompressible flow? If so, are there
any limitations on this equation?
W4.3 Consider a CD (compact disc) rotating at angular velocity Ω. Does it have vorticity in the sense of this chapter?
If so, how much vorticity?
Comprehensive Problems 291
W4.4 How much acceleration can fluids endure? Are fluids
like astronauts, who feel that 5g is severe? Perhaps use
the flow pattern of Example 4.8, at r = R, to make some
estimates of fluid acceleration magnitudes.
W4.5 State the conditions (there are more than one) under
which the analysis of temperature distribution in a flow
field can be completely uncoupled, so that a separate
analysis for velocity and pressure is possible. Can we do
this for both laminar and turbulent flow?
W4.6 Consider liquid flow over a dam or weir. How might the
boundary conditions and the flow pattern change when
we compare water flow over a large prototype to SAE 30
oil flow over a tiny scale model?
W4.7 What is the difference between the stream function ψ and
our method of finding the streamlines from Sec. 1.11?
Or are they essentially the same?
W4.8 Under what conditions do both the stream function ψ and
the velocity potential ϕ exist for a flow field? When does
one exist but not the other?
W4.9 How might the remarkable three-dimensional Taylor
­instability of Fig. 4.14 be predicted? Discuss a general
­procedure for examining the stability of a given flow
­pattern.
W4.10 Consider an irrotational, incompressible, axisymmetric
(∂/∂θ = 0) flow in (r, z) coordinates. Does a stream function exist? If so, does it satisfy Laplace’s equation? Are
lines of constant ψ equal to the flow streamlines? Does a
velocity potential exist? If so, does it satisfy Laplace’s
equation? Are lines of constant ϕ everywhere perpendicular to the ψ lines?
Fundamentals of Engineering Exam Problems
This chapter is not a favorite of the people who prepare the FE
Exam. Probably not a single problem from this chapter will appear on the exam, but if some did, they might be like these.
FE4.1 Given the steady, incompressible velocity distribution
V = 3xi + Cyj + 0k, where C is a constant, if conservation of mass is satisfied, the value of C should be
(a) 3, (b) 3/2, (c) 0, (d ) –3/2, (e) –3
FE4.2 Given the steady velocity distribution V = 3xi + 0j +
Cyk, where C is a constant, if the flow is irrotational, the
value of C should be
(a) 3, (b) 3/2, (c) 0, (d ) –3/2, (e) –3
FE4.3 Given the steady, incompressible velocity distribution
V = 3xi + Cyj + 0k, where C is a constant, the shear
stress τxy at the point (x, y, z) is given by
(a) 3µ, (b) (3x + Cy)µ, (c) 0, (d ) Cµ, (e) (3 + C)µ
FE4.4 Given the steady, incompressible velocity distribution
u = Ax, υ = By, and w = Cxy, where (A, B, C) are constants. This flow satisfies the equation of continuity if
A equals
(a) B, (b) B + C, (c) B – C, (d ) –B, (e) –(B + C )
FE4.5 For the velocity field in Prob. FE4.4, the convective
­acceleration in the x direction is
(a) Ax2, (b) A2x, (c) B2y, (d ) By2, (e) Cx2y
FE4.6 If, for laminar flow in a smooth, straight tube, the tube
­diameter and length both double, while everything else
­remains the same, the volume flow rate will ­increase by
a factor of
(a) 2, (b) 4, (c) 8, (d ) 12, (e) 16
Comprehensive Problems
In a certain medical application, water at room temperature and pressure flows through a rectangular channel of
length L = 10 cm, width s = 1.0 cm, and gap thickness b
= 0.30 mm as in Fig. C4.1. The volume flow rate is sinusoidal with amplitude Q̂ = 0.50 mL/s and frequency f
= 20 Hz, i.e., Q = Q̂ sin (2πft).
(a) Calculate the maximum Reynolds number (Re =
Vb/υ) based on maximum average velocity and gap
thickness. Channel flow like this remains laminar for Re
less than about 2000. If Re is greater than about 2000,
the flow will be turbulent. Is this flow laminar or turbulent? (b) In this problem, the frequency is low enough
C4.1
that at any given time, the flow can be solved as if it
were steady at the given flow rate. (This is called a
quasi-steady assumption.) At any arbitrary instant of
time, find an expression for streamwise velocity u as a
function of y, µ, dp/dx, and b, where dp/dx is the pressure gradient required to push the flow through the
channel at volume flow rate Q. In addition, ­estimate the
maximum magnitude of velocity component u. (c) At
any instant of time, find a relationship between volume
flow rate Q and pressure gradient dp/dx. Your a­ nswer
should be given as an expression for Q as a function of
dp/dx, s, b, and viscosity µ. (d ) Estimate the wall shear
292
Chapter 4 Differential Relations for Fluid Flow
stress at the outer film edge, derive a formula for (a) υ(x),
(b) the average velocity Vavg in the film, and (c) the velocity Vc for which there is no net flow either up or down.
(d ) Sketch υ(x) for case (c).
stress, τw as a function of Q̂, f, µ, b, s, and time (t).
(e) ­Finally, for the numbers given in the problem statement, estimate the amplitude of the wall shear stress,
τ̂w , in N/m2.
h ≈ constant
L
y
s
x
y, v
z
V
x, u
ρ, µ
Q
b
C4.1
C4.2
C4.2
A belt moves upward at velocity V, dragging a film of
­viscous liquid of thickness h, as in Fig. C4.2. Near the
belt, the film moves upward due to no slip. At its outer
edge, the film moves downward due to gravity. Assuming that the only nonzero velocity is υ(x), with zero shear
Belt
References
1.
2.
3.
4.
5.
6.
J. D. Anderson, Computational Fluid Dynamics: An
Introduction, 3d ed., Springer, New York, 2010.
C. E. Brennen, Fundamentals of Multiphase Flow,
­Cambridge University Press, New York, 2009. See also
URL <http://caltechbook.library.caltech.edu/51/01/
multiph.htm>
D. Zwillinger, CRC Standard Mathematical Tables and
­Formulae, 32d ed., CRC Press Inc., Cleveland, Ohio,
2011.
H. Schlichting and K. Gersten, Boundary Layer Theory,
8th ed., Springer, New York, 2000.
F. M. White, Viscous Fluid Flow, 3d ed., McGraw-Hill,
New York, 2005.
E. B. Tadmor, R. E. Miller, and R. S. Elliott, Continuum
­Mechanics and Thermodynamics, Cambridge University
Press, New York, 2012.
7.
8.
9.
10.
11.
12.
13.
J. P. Holman, Heat Transfer, 10th ed., McGraw-Hill, New
York, 2009.
W. M. Kays and M. E. Crawford, Convective Heat and
Mass Transfer, 4th ed., McGraw-Hill, New York, 2004.
G. K. Batchelor, An Introduction to Fluid Dynamics,
­Cambridge University Press, Cambridge, England, 1967.
L. Prandtl and O. G. Tietjens, Fundamentals of Hydro- and
Aeromechanics, Dover, New York, 1957.
D. Fleisch, A Student’s Guide to Vectors and Tensors,
­Cambridge University Press, New York, 2011.
O. Gonzalez and A. M. Stuart, A First Course in Continuum Mechanics, Cambridge University Press, New York,
2008.
D. A. Danielson, Vectors and Tensors in Engineering and
Physics, 2d ed., Westview (Perseus) Press, Boulder, CO,
2003.
References 293
14.
R. I. Tanner, Engineering Rheology, 2d ed., Oxford
­University Press, New York, 2000.
15. H. Lamb, Hydrodynamics, 6th ed., Dover, New York, 1945.
16. J. P. Davin, Tribology for Engineers: A Practical Guide,
Woodhead Publishing, Philadelphia, 2011.
17.
G. I. Taylor, “Stability of a Viscous Liquid Contained
­between Two Rotating Cylinders,” Philos. Trans. Roy. Soc.
London Ser. A, vol. 223, 1923, pp. 289–343.
A full-scale NASA parachute, which helped lower the vehicle Curiosity to the Mars surface
in 2012, was tested in the world’s largest wind tunnel, at NASA Ames Research Center,
Moffett Field, California. It is the largest disc-gap-band parachute [36] ever built, with a
diameter of 51 feet. In the Mars atmosphere it will generate up to 65,000 lbf of drag,
which leads to a problem assignment in Chap. 7. [Image from JPL-Caltech/NASA.]
294
Chapter 5
Dimensional Analysis
and Similarity
Motivation. In this chapter we discuss the planning, presentation, and interpreta-
tion of experimental data. We shall try to convince you that such data are best
presented in dimensionless form. Experiments that might result in tables of output,
or even multiple volumes of tables, might be reduced to a single set of curves—or
even a single curve—when suitably nondimensionalized. The technique for doing
this is dimensional analysis. It is also effective in theoretical studies.
Chapter 3 presented large-scale control volume balances of mass, momentum,
and energy, which led to global results: mass flow, force, torque, total work done,
or heat transfer. Chapter 4 presented infinitesimal balances that led to the basic
partial differential equations of fluid flow and some particular solutions for both
inviscid and viscous (laminar) flow. These straight analytical techniques are limited to simple geometries and uniform boundary conditions. Only a fraction of
engineering flow problems can be solved by direct analytical formulas.
Most practical fluid flow problems are too complex, both geometrically and
physically, to be solved analytically. They must be tested by experiment or approximated by computational fluid dynamics (CFD) [2]. These results are typically
reported as experimental or numerical data points and smoothed curves. Such
data have much more generality if they are expressed in compact, economic form.
This is the motivation for dimensional analysis. The technique is a mainstay of
fluid mechanics and is also widely used in all engineering fields plus the physical, biological, medical, and social sciences. The present chapter shows how
dimensional analysis improves the presentation of both data and theory.
5.1 Introduction
Basically, dimensional analysis is a method for reducing the number and complexity of experimental variables that affect a given physical phenomenon, by using
a sort of compacting technique. If a phenomenon depends on n dimensional
295
296
Chapter 5 Dimensional Analysis and Similarity
variables, dimensional analysis will reduce the problem to only k dimensionless
variables, where the reduction n − k = 1, 2, 3, or 4, depending on the problem
complexity. Generally n − k equals the number of different dimensions (sometimes called basic or primary or fundamental dimensions) that govern the problem. In fluid mechanics, the four basic dimensions are usually taken to be mass
M, length L, time T, and temperature Θ, or an MLTΘ system for short. Alternatively, one uses an FLTΘ system, with force F replacing mass.
Although its purpose is to reduce variables and group them in dimensionless
form, dimensional analysis has several side benefits. The first is enormous savings
in time and money. Suppose one knew that the force F on a particular body shape
immersed in a stream of fluid depended only on the body length L, stream velocity V, fluid density ρ, and fluid viscosity µ; that is,
F = f(L, V, ρ, μ)
(5.1)
Suppose further that the geometry and flow conditions are so complicated that
our integral theories (Chap. 3) and differential equations (Chap. 4) fail to yield
the solution for the force. Then we must find the function f(L, V, ρ, µ) experimentally or numerically.
To conduct the experiments systematically, it would be necessary to change
the variables one at time while holding all others constant. For example, we can
make several bodies in different lengths and conduct the tests under the same
setup so that we can examine the effect of the body length on the force in
Eq. (5.1). Generally speaking, it takes about 10 points to define a curve. This
series of tests would yield data that could eventually plot the variations of each
variable on the force in four different curves. Although it sounds logical, this
experimental approach to finding out the relation between the force and various
factors can be hardly implemented. First of all, the tests are costly in time and
money. In addition, one of the experiments in this case needs to vary 10 different
fluid densities while holding the viscosity constant. Finally, it remains uncertain
that how we can combine the curves obtained individually from each set of independent variable to generate the desired functional relationship among the dependent variable force and other independent variables in Eq. (5.1) for the application
to any similar external flow problem. However, with dimensional analysis, we can
immediately reduce Eq. (5.1) to the equivalent form
F
2 2
or
= g(
ρVL μ )
ρV L
CF = g(Re)
(5.2)
That is, the dimensionless force coefficient F/(ρV2L2) is a function only of the
dimensionless Reynolds number ρVL/µ. We shall learn exactly how to make this
reduction in Secs. 5.2 and 5.3. Equation (5.2) will be useful in Chap. 7.
Note that Eq. (5.2) is just an example, not the full story, of forces caused by fluid
flows. Some fluid forces have a very weak or negligible Reynolds number dependence in wide regions (Fig. 5.2a). Other groups may also be important. The force
coefficient may depend, in high-speed gas flow, on the Mach number, Ma = V/a,
where a is the speed of sound. In free-surface flows, such as ship drag, CF may
5.1 Introduction 297
depend upon Froude number, Fr = V2/(gL), where g is the acceleration of gravity. In
turbulent flow, force may depend upon the roughness ratio, ϵ/L, where ϵ is the roughness height of the surface.
The function g is different mathematically from the original function f, but it
contains all the same information. Nothing is lost in a dimensional analysis. And
think of the savings: We can establish g by running the experiment for only
10 values of the single variable called the Reynolds number. We do not have to
vary L, V, ρ, or µ separately but only the grouping ρVL/µ. This we do merely by
varying velocity V in, say, a wind tunnel or drop test or water channel, and there
is no need to build 10 different bodies or find 100 different fluids with 10 densities and 10 viscosities. The cost is now about $1000, maybe less.
A second side benefit of dimensional analysis is that it helps our thinking and
planning for an experiment or theory. It suggests dimensionless ways of writing
equations before we spend money on computer analysis to find solutions. It suggests variables that can be discarded; sometimes dimensional analysis will immediately reject variables, and at other times it groups them off to the side, where a
few simple tests will show them to be unimportant. Finally, dimensional analysis
will often give a great deal of insight into the form of the physical relationship
we are trying to study.
A third benefit is that dimensional analysis provides scaling laws that can
convert data from a cheap, small model to design information for an expensive,
large prototype. We do not build a million-dollar airplane and see whether it has
enough lift force. We measure the lift on a small model and use a scaling law to
predict the lift on the full-scale prototype airplane. There are rules we shall
explain for finding scaling laws. When the scaling law is valid, we say that a
condition of similarity exists between the model and the prototype. In the simple
case of Eq. (5.1), similarity is achieved if the Reynolds number is the same for
the model and prototype because the function g then requires the force coefficient
to be the same also:
If Rem = Rep then CFm = CFp (5.3)
where subscripts m and p mean model and prototype, respectively. From the
definition of force coefficient, this means that
Fp
Fm
=
ρp Vp 2 Lp 2
ρm ( Vm ) ( Lm )
(5.4)
for data taken where ρpVpLp/µp = ρmVmLm/µm. Equation (5.4) is a scaling law: If
you measure the model force at the model Reynolds number, the prototype force
at the same Reynolds number equals the model force times the density ratio times
the velocity ratio squared times the length ratio squared. We shall give more
examples later.
Do you understand these introductory explanations? Be careful; learning dimensional analysis is like learning to play tennis: There are levels of the game. We can
establish some ground rules and do some fairly good work in this brief chapter, but
dimensional analysis in the broad view has many subtleties and nuances that only
time, practice, and maturity enable you to master. Although dimensional analysis
298
Chapter 5 Dimensional Analysis and Similarity
has a firm physical and mathematical foundation, considerable art and skill are
needed to use it effectively.
EXAMPLE 5.1
A copepod is a water crustacean approximately 1 mm in diameter. We want to know
the drag force on the copepod when it moves slowly in fresh water. A scale model 100
times larger is made and tested in glycerin at V = 30 cm/s. The measured drag on the
model is 1.3 N. For similar conditions, what are the velocity and drag of the actual
copepod in water? Assume that Eq. (5.2) applies and the temperature is 20°C.
Solution
∙ Property values: From Table A.3, the densities and viscosities at 20°C are
Water (prototype):
µp = 0.001 kg/(m-s)
ρp = 998 kg/m3
Glycerin (model):
µm = 1.5 kg/(m-s)
ρm = 1263 kg/m3
∙ Assumptions: Equation (5.2) is appropriate and similarity is achieved; that is, the
model and prototype have the same Reynolds number and, therefore, the same force
coefficient.
∙ Approach: The length scales are Lm = 100 mm and Lp = 1 mm. Calculate the
Reynolds number and force coefficient of the model and set them equal to prototype
values:
Rem =
(998 kg/m3 )Vp (0.001 m)
ρmVm Lm (1263 kg/m3 ) (0.3 m/s) (0.1 m)
=
= 25.3 = Rep =
μm
1.5 kg/(m-s)
0.001 kg/(m-s)
Solve for Vp = 0.0253 m/s = 2.53 cm/s
Ans.
In like manner, using the prototype velocity just found, equate the force coefficients:
CFm =
Fm
ρmVm2 L2m
= CFp =
=
1.3 N
= 1.14
(1263 kg/m ) (0.3 m/s) 2 (0.1 m) 2
3
Fp
3
(998 kg/m ) (0.0253 m/s) 2 (0.001 m) 2
Solve for Fp = 7.3E−7 N
Ans.
∙ Comments: Assuming we modeled the Reynolds number correctly, the model test is
a very good idea, as it would obviously be difficult to measure such a tiny copepod
drag force.
Historically, the first person to write extensively about units and dimensional
reasoning in physical relations was Euler in 1765. Euler’s ideas were far ahead
of his time, as were those of Joseph Fourier, whose 1822 book Analytical Theory
of Heat outlined what is now called the principle of dimensional homogeneity
and even developed some similarity rules for heat flow. There were no further
significant advances until Lord Rayleigh’s book in 1877, Theory of Sound, which
proposed a “method of dimensions” and gave several examples of dimensional
5.2 The Principle of Dimensional Homogeneity 299
analysis. The final breakthrough that established the method as we know it today
is generally credited to E. Buckingham in 1914 [1], whose paper outlined what
is now called the Buckingham Pi Theorem for describing dimensionless parameters (see Sec. 5.3). However, it is now known that a Frenchman, A. Vaschy, in
1892 and a Russian, D. Riabouchinsky, in 1911 had independently published
papers reporting results equivalent to the pi theorem. Following Buckingham’s
paper, P. W. Bridgman published a classic book in 1922 [3], outlining the general
theory of dimensional analysis.
Dimensional analysis is so valuable and subtle, with both skill and art
involved, that it has spawned a wide variety of textbooks and treatises. The
writer is aware of more than 30 books on the subject, of which his engineering
favorites are listed here [3–10]. Dimensional analysis is not confined to fluid
mechanics, or even to engineering. S
­ pecialized books have been published on
the application of dimensional analysis to metrology [11], astrophysics [12],
economics [13], chemistry [14], hydrology [15], medications [16], clinical
medicine [17], chemical processing pilot plants [18], social sciences [19], biomedical sciences [20], pharmacy [21], fractal geometry [22], and even the
growth of plants [23]. Clearly this is a subject well worth learning for many
career paths.
5.2 The Principle of Dimensional Homogeneity
In making the remarkable jump from the five-variable Eq. (5.1) to the two-variable Eq. (5.2), we were exploiting a rule that is almost a self-evident axiom in
physics. This rule, the principle of dimensional homogeneity (PDH), can be stated
as follows:
If an equation truly expresses a proper relationship between variables in a physical
process, it will be dimensionally homogeneous; that is, each of its additive terms will
have the same dimensions.
All the equations that are derived from the theory of mechanics are of this form.
For example, consider the relation that expresses the displacement of a falling
body:
S = S0 + V0t + 12gt2
(5.5)
Each term in this equation is a displacement, or length, and has dimensions {L}.
The equation is dimensionally homogeneous. Note also that any consistent set of
units can be used to calculate a result.
Consider Bernoulli’s equation for incompressible flow:
p 1 2
+ V + gz = const
ρ 2
(5.6)
Each term, including the constant, has dimensions of velocity squared, or {L2T −2}.
The equation is dimensionally homogeneous and gives proper results for any
consistent set of units.
300
Chapter 5 Dimensional Analysis and Similarity
Students count on dimensional homogeneity and use it to check themselves
when they cannot quite remember an equation during an exam. For example,
which is it:
S = 12gt2? or S = 12g2t?
(5.7)
By checking the dimensions, we reject the second form and back up our faulty
memory. We are exploiting the principle of dimensional homogeneity, and this
chapter simply exploits it further.
Variables and Constants
Equations (5.5) and (5.6) also illustrate some other factors that often enter into a
dimensional analysis:
Dimensional variables are the quantities that actually vary during a given case
and would be plotted against each other to show the data. In Eq. (5.5), they
are S and t; in Eq. (5.6) they are p, V, and z. All have dimensions, and all
can be nondimensionalized as a dimensional analysis technique.
Dimensional constants may vary from case to case but are held constant
during a given run. In Eq. (5.5) they are S0, V0, and g, and in Eq. (5.6)
they are ρ, g, and C. They all have dimensions and conceivably could be
nondimensionalized, but they are normally used to help nondimensionalize the variables in the problem.
Pure constants have no dimensions and never did. They arise from mathematical manipulations. In both Eqs. (5.5) and (5.6) they are 12 and the exponent 2, both of which came from an integration: et dt = 12t2, eV dV = 21V2.
Other common dimensionless constants are π and e. Also, the argument of any mathematical function, such as ln, exp, cos, or J0, is
dimensionless.
Angles and revolutions are dimensionless. The preferred unit for an angle is
the radian, which makes it clear that an angle is a ratio. In like manner, a
revolution is 2π radians.
Counting numbers are dimensionless. For example, if we triple the energy E
to 3E, the coefficient 3 is dimensionless.
Note that integration and differentiation of an equation may change the
dimensions but not the homogeneity of the equation. For example, integrate or
differentiate Eq. (5.5):
∫ S dt = S t +
0
1
2
2 V0t
+ 16gt 3
dS
= V0 + gt
dt
(5.8a)
(5.8b)
In the integrated form (5.8a) every term has dimensions of {LT}, while in the
derivative form (5.8b) every term is a velocity {LT−1}.
Finally, some physical variables are naturally dimensionless by virtue of their
definition as ratios of dimensional quantities. Some examples are strain (change
5.3 The Pi Theorem 301
in length per unit length), Poisson’s ratio (ratio of transverse strain to longitudinal
strain), and specific gravity (ratio of density to standard water density).
The motive behind dimensional analysis is that any dimensionally homogeneous equation can be written in an entirely equivalent nondimensional form that
is more compact. Usually there are multiple methods of presenting one’s dimensionless data or theory.
5.3 The Pi Theorem
There are several methods of reducing a number of dimensional variables into a
smaller number of dimensionless groups. The first scheme given here was proposed in 1914 by Buckingham [1] and is now called the Buckingham Pi Theorem.
The name pi comes from the mathematical notation Π, meaning a product of
variables. The dimensionless groups found from the theorem are power products
denoted by Π1, Π2, Π3, etc. The method allows the pi groups to be found in
sequential order without resorting to free exponents.
The first part of the pi theorem explains what reduction in variables to expect:
If a physical process satisfies the PDH and involves n dimensional variables, it can be
reduced to a relation between k dimensionless variables or Πs. The reduction j = n − k
equals the maximum number of variables that do not form a pi among themselves and
is always less than or equal to the number of dimensions describing the variables.
Take the specific case of force on an immersed body: Eq. (5.1) contains five
variables F, L, V, ρ, and µ described by three dimensions {MLT}. Thus n = 5
and j ≤ 3. Therefore it is a good guess that we can reduce the problem to k pi
groups, with k = n − j ≥ 5 − 3 = 2. And this is exactly what we obtained: two
dimensionless ­variables Π1 = CF and Π2 = Re. On rare occasions it may take
more pi groups than this minimum (see Example 5.5).
The second part of the theorem shows how to find the pi groups one at a time:
Find the reduction j, then select j scaling variables that do not form a pi among themselves.1 Each desired pi group will be a power product of these j variables plus one
additional variable, which is assigned any convenient nonzero exponent. Each pi group
thus found is independent.
Several methods can be used to form the dimensionless products or pi groups. A
well-adopted method called the repeating variable method of dimensional analysis will be described in detail first followed by a few more examples. We will
briefly introduce an alternate method by Ispen [5] at the end. For better explanation, it will be helpful for beginners to break the repeating variable method down
into a series of distinct steps, as shown in Example 5.2 with additional discussions. In this way, with a little practice you will be able to complete a dimensional
analysis for a specific problem of your own.
1
Make a clever choice here because all pi groups will contain these j variables in various
groupings.
302
Chapter 5 Dimensional Analysis and Similarity
EXAMPLE 5.2
Repeat the development of Eq. (5.2) from Eq. (5.1), using the pi theorem.
Solution
Step 1
Write the function with all the variables that are involved in the problem, and count the
variables:
F = f(L, V, ρ, μ)
there are five variables (n = 5)
The foundation of the dimensional analysis method rests on two assumptions: that
(1) the proposed physical relation is dimensionally homogeneous as we demonstrated
in Sec. 5.2, and (2) all the relevant variables have been included in the proposed
relation.
This step is in fact the most difficult one in real engineering practice, although
such a relation is provided in this example and in all of the end-of-chapter problems.
If a relevant variable is missing, dimensional analysis will fail, resulting in either
algebraic difficulties or, worse, a dimensionless formulation that does not resolve the
process. Typically the variables will include (1) geometric variables of the system,
such as the length of a body, the diameter of a pipe; (2) variables of fluid properties
like density, viscosity; (3) variables representing external effects, such as drag force
of an immersed body, pressure drop along a pipeline. On the other hand, we wish to
keep the number of variables to a minimum, so that we can minimize the amount
laboratory work. Therefore, the determination of the variables must be accomplished
by one’s understanding of the problem and the physical laws that govern the phenomenon. A primary difficulty is that it is not easy to determine which variables to include
or exclude.
Step 2
List dimensions of each variable. From Table 5.1
F
−2
{MLT }
L
{L}
V
−1
{LT }
ρ
μ
−3
{ML }
−1
{ML T −1}
Step 3
Find the reduction j and determine the number of pi groups to be formed. No variable
contains the dimension Θ, and so j is less than or equal to 3 (MLT). We inspect the list
and see that L, U, and ρ cannot form a pi group because only ρ contains mass and only
U contains time. Therefore j does equal 3, and n − j = 5 − 3 = 2 = k. The pi theorem
guarantees for this problem that there will be exactly two independent ­dimensionless
groups.
Step 4
Select j scaling variables that do not form a pi product.
As a rule, each selected scaling variable must be dimensionally independent of the
others. This means that they must not form a dimensionless group among themselves,
5.3 The Pi Theorem 303
but adding one more variable will form a dimensionless quantity. For example, test
powers of ρ, V, and L:
ρaVbLc = (ML−3 ) a (L/T) b (L) c = M0L0T0 only if a = 0, b = 0, c = 0
In this case, we can see why this is so: Only ρ contains the dimension {M}, and only
V contains the dimension {T}, so no cancellation is possible. If, now, we add μ to the
scaling group, we will obtain the Reynolds number. If we add F to the group, we form
the force coefficient.
However, if variables μ, V, and L or μ, ρ, and L are selected, they still obey the
above rule. In other words, selection of the scaling variables is not unique. Clearly there
is ambiguity in these choices, something that often vexes the beginning experimenter.
But the ambiguity is deliberate. Its purpose is to show a particular effect, and the choice
is yours. The followings are two guidelines for selecting scaling variables:
1. Do not select output variables for your scaling variables. In Eq. (5.1), certainly
do not select the dependent variable F, which you wish to isolate for your plot.
Nor was μ selected, for we wished to plot force versus viscosity.
2. If convenient, select simple and popular, not obscure, scaling variables because they
will appear in all of your dimensionless groups. Select density, not surface tension.
Select body length, not volume. Select stream velocity, not speed of sound.
In another occasion, suppose we wish to study drag force versus velocity. Then we
would not use V as a scaling variable in Eq. (5.1). We would use (ρ, μ, L) instead, and
the final dimensionless function would become
CF′ =
ρF
μ
2
= f(Re)
Re =
ρVL
μ
In plotting these data, we would not be able to discern the effect of ρ or μ, since they
appear in both dimensionless groups. The grouping CF′ again would mean dimensionless
force, and Re is now interpreted as either dimensionless velocity or size as we notice
that L as a scaling variable did not appear in the drag coefficient in this case. The plot
would be quite different compared to Eq. (5.2), although it contains exactly the same
information. The development of dimensionless groups such as CF′ and Re from the
initial variables is the subject of the pi theorem.
According to the rule and guidelines, the group L, V, ρ we found in step 3 will do
fine. They will be used as the scaling variables in step 5 below.
Step 5
Combine L, U, ρ with one additional variable, in sequence, to find the two pi products.
First add force to find Π1. You may select any exponent on this additional term as you
please, to place it in the numerator or denominator to any power. Since F is the output,
or dependent, variable, we select it to appear to the first power in the numerator:
Π 1 = LaUbρcF = (L) a (LT −1 ) b (ML−3 ) c (MLT−2 ) = M0L0T0
Equate exponents:
Length:
a + b − 3c + 1 = 0
Mass:
Time:
c+1=0
−b
−2 = 0
304
Chapter 5 Dimensional Analysis and Similarity
We can solve explicitly for
a = −2
b = −2
Π 1 = L−2U−2ρ−1F =
Therefore
c = −1
F
= CF ρU2L2
Ans.
This is exactly the right pi group as in Eq. (5.2). By varying the exponent on F, we
could have found other equivalent groups such as ULρ1/2/F1/2.
Finally, add viscosity to L, U, and ρ to find Π2. Select any power you like for
viscosity. By hindsight and custom, we select the power −1 to place it in the
denominator:
Π 2 = LaUbρcμ−1 = La (LT −1 ) b (ML−3 ) c (ML−1T −1 ) −1 = M0L0T0
Equate exponents:
Length:
a + b − 3c + 1 = 0
Mass:
Time:
c−1=0
−b
+1=0
from which we find
a=b=c=1
Therefore
Π 2 = L1U1ρ1μ−1 =
ρUL
= Re
μ
Ans.
There is also ambiguity in selecting the power of the nonscaling variable to find pi
products as long as it is nonzero. In general, we can leave it as +1 because it is the
simplest power form. But we selected −1 for μ in the process to form the second pi
product Re in above example. What pi product will result in if +1 is selected as the
power for μ?
Step 6
Express the final form as a relationship among the pi groups. The pi theorem guarantees
that the functional relationship must be of the equivalent form
ρVL
F
= g(
2 2
μ )
ρV L
Ans.
which is exactly Eq. (5.2).
The six steps demonstrated in Example 5.2 are summarized below to outline
a recommended procedure for determining pi groups:
1. List and count the n variables involved in the problem. If any important
variables are missing, dimensional analysis will fail.
2. List the dimensions of each variable according to {MLTΘ} or {FLTΘ}. A
list is given in Table 5.1.
5.3 The Pi Theorem 305
Table 5.1 Dimensions of Fluid-Mechanics Properties
Dimensions
Quantity
Length
Area
Volume
Velocity
Acceleration
Speed of sound
Volume flow
Mass flow
Pressure, stress
Strain rate
Angle
Angular velocity
Viscosity
Kinematic viscosity
Surface tension
Force
Moment, torque
Power
Work, energy
Density
Temperature
Specific heat
Specific weight
Thermal conductivity
Thermal expansion coefficient
Symbol
L
A
𝒱
V
dV/dt
a
Q
ṁ
p, σ, τ
ε̇
θ
ω, Ω
µ
ν
Υ
F
M
P
W, E
ρ
T
cp, cυ
γ
k
β
MLTΘ
L
L2
L3
LT −1
LT −2
LT −1
L3T −1
MT −1
ML−1T −2
T −1
None
T −1
ML−1T −1
L2T −1
MT −2
MLT −2
ML2T −2
ML2T –3
ML2T −2
ML−3
Θ
L2T −2Θ−1
ML–2T −2
MLT –3Θ−1
Θ−1
FLTΘ
L
L2
L3
LT −1
LT −2
LT −1
L3T −1
FTL−1
FL−2
T −1
None
T −1
FTL−2
L2T −1
FL−1
F
FL
FLT −1
FL
FT2L−4
Θ
L2T −2Θ−1
FL−3
FT −1Θ−1
Θ−1
3. Find the reduction j and determine the number of pi groups to be formed.
Initially guess j equal to the number of different dimensions present, and
look for j variables that do not form a pi product. If no luck, reduce j by
1 and look again. With practice, you will find j ­rapidly.
4. Select j scaling parameters that do not form a pi product. Make sure they
please you and have some generality if possible, because they will then
appear in every one of your pi groups. Pick density or velocity or length. Do
not pick surface tension, for example, or you will form six different independent Weber-number parameters and thoroughly annoy your colleagues.
5. Add one additional variable to your j repeating variables, and form a power
product. Algebraically find the exponents that make the product dimensionless. Try to arrange for your output or dependent variables (force, pressure
drop, torque, power) to appear in the numerator, and your plots will look
better. Do this sequentially, adding one new variable each time, and you will
find all n − j = k desired pi products.
6. Write the final dimensionless function, and check the terms to make sure
all pi groups are dimensionless.
306
Chapter 5 Dimensional Analysis and Similarity
EXAMPLE 5.3
The power input P to a centrifugal pump is a function of the volume flow Q, impeller
diameter D, rotational rate Ω, and the density ρ and viscosity μ of the fluid:
P = f(Q, D, Ω, ρ, μ)
Rewrite this as a dimensionless relationship. Hint: Use Ω, ρ, and D as repeating
­variables. We will revisit this problem in Chap. 11.
Solution
Step 1
Count the variables. There are six (don’t forget the one on the left, P).
Step 2
List the dimensions of each variable from Table 5.1. Use the {FLTΘ} system:
P
Q
−1
{FLT }
3
Ω
D
−1
{L T }
ρ
−1
{L}
2 −4
{T }
{FT L }
µ
{FTL−2}
Step 3
Find j. Lucky us, we were told to use (Ω, ρ, D) as repeating variables, so surely
j = 3, the number of dimensions (FLT)? Check that these three do not form a pi
group:
ΩaρbDc = (T −1 ) a (FT 2L−4 ) b (L) c = F0L0T0 only if
a = 0, b = 0, c = 0
Yes, j = 3. This was not as obvious as the scaling group (L, U, ρ) in Example 5.2, but
it is true. We now know, from the theorem, that adding one more variable will indeed
form a pi group.
Step 4a
Combine (Ω, ρ, D) with power P to find the first pi group:
Π 1 = ΩaρbDcP = (T −1 ) a (FT 2L−4 ) b (L) c (FLT −1 ) = F0L0T0
Equate exponents:
Force:
Length:
Time:
b
+1 = 0
−4b + c + 1 = 0
−a + 2b
−1 = 0
Solve algebraically to obtain a = −3, b = −1, and c = −5. This first pi group, the
output dimensionless variable, is called the power coefficient of a pump, CP:
Π 1 = Ω−3ρ−1D−5P =
P
ρΩ3D5
= CP
5.3 The Pi Theorem 307
Step 4b
Combine (Ω, ρ, D) with flow rate Q to find the second pi group:
Π 2 = ΩaρbDcQ = (T −1 ) a (FT 2L−4 ) b (L) c (L3T −1 ) = F0L0T0
After equating exponents, we now find a = −1, b = 0, and c = −3. This second pi
group is called the flow coefficient of a pump, CQ:
Q
Π 2 = Ω−1ρ0D−3Q =
ΩD3
= CQ
Step 4c
Combine (Ω, ρ, D) with viscosity μ to find the third and last pi group:
Π 3 = ΩaρbDcμ = (T −1 ) a (FT 2L−4 ) b (L) c (FTL−2 ) = F0L0T0
This time, a = −1, b = −1, and c = −2; or Π3 = µ/(ρΩD2), a sort of Reynolds
number.
Step 5
The original relation between six variables is now reduced to three dimensionless groups:
P
3
ρΩ D
5
=f(
Q
ΩD
3
,
ρΩD2 )
μ
Ans.
Comment: These three are the classical coefficients used to correlate pump power
in Chap. 11.
EXAMPLE 5.4
At low velocities (laminar flow), the volume flow Q through a small-bore tube is a
function only of the tube radius R, the fluid viscosity µ, and the pressure drop per unit
tube length dp/dx. Using the pi theorem, find an appropriate dimensionless relationship.
Solution
Write the given relation and count variables:
dp
Q = f (R, μ, ) four variables (n = 4)
dx
Make a list of the dimensions of these variables from Table 5.1 using the {MLT}
system:
Q
3
−1
{L T }
R
μ
{L}
−1
dp/dx
−1
{ML T }
{ML−2T −2}
There are three primary dimensions (M, L, T ), hence j ≤ 3. By trial and error we determine
that R, μ, and dp/dx cannot be combined into a pi group. Then j = 3, and n − j = 4 − 3 = 1.
308
Chapter 5 Dimensional Analysis and Similarity
There is only one pi group, which we find by combining Q in a power product with
the other three:
Π 1 = Ra μb (
dp c 1
Q = (L) a (ML−1T −1 ) b (ML−2T −2 ) c (L3T −1 )
dx )
= M0L0T 0
Equate exponents:
Mass:
b + c =0
Length:
a − b − 2c + 3 = 0
Time:
−b − 2c − 1 = 0
Solving simultaneously, we obtain a = −4, b = 1, and c = −1. Then
Π 1 = R−4μ1(
or
Π1 =
dp −1
Q
dx )
Qμ
4
R (dp∕dx)
= const
Ans.
Since there is only one pi group, it must equal a dimensionless constant. This is as far
as dimensional analysis can take us. The laminar flow theory of Sec. 4.10 shows that
the value of the constant is −π8 . This result is also useful in Chap. 6.
EXAMPLE 5.5
Assume that the tip deflection δ of a cantilever beam is a function of the tip load P,
beam length L, area moment of inertia I, and material modulus of elasticity E; that is,
δ = f(P, L, I, E). Rewrite this function in dimensionless form, and comment on its
complexity and the peculiar value of j.
Solution
List the variables and their dimensions:
δ
{L}
P
L
−2
{MLT }
{L}
I
4
{L }
E
−1
{ML T −2}
There are five variables (n = 5) and three primary dimensions (M, L, T ), hence j ≤ 3.
But try as we may, we cannot find any combination of three variables that does not
form a pi group. This is because {M} and {T} occur only in P and E and only in the
same form, {MT−2}. Thus we have encountered a special case of j = 2, which is less
than the number of dimensions (M, L, T). To gain more insight into this peculiarity,
you should rework the problem, using the (F, L, T) system of dimensions. You will
find that only {F} and {L} occur in these variables, hence j = 2.
With j = 2, we select L and E as two variables that cannot form a pi group and
then add other variables to form the three desired pis:
Π 1 = LaEbI1 = (L) a (ML−1T −2 ) b (L4 ) = M0L0T0
5.3 The Pi Theorem 309
from which, after equating exponents, we find that a = −4, b = 0, or Π1 = I/L4. Then
Π 2 = LaEbP1 = (L) a (ML−1T −2 ) b (MLT−2 ) = M0L0T0
from which we find a = −2, b = −1, or Π2 = P/(EL2), and
Π 3 = LaEbδ1 = (L) a (ML−1T −2 ) b (L) = M0L0T0
from which a = −1, b = 0, or Π3 = δ/L. The proper dimensionless function is Π3 =
f(Π2, Π1), or
δ
P I
= f ( 2 , 4 )
L
EL L
Ans. (1)
This is a complex three-variable function, but dimensional analysis alone can take us
no further.
Comments: We can “improve” Eq. (1) by taking advantage of some physical reasoning, as Langhaar points out [4, p. 91]. For small elastic deflections, δ is proportional
to load P and inversely proportional to moment of inertia I. Since P and I occur separately in Eq. (1), this means that Π3 must be proportional to Π2 and inversely proportional to Π1. Thus, for these conditions,
δ
P L4
= (const)
L
EL2 I
PL3
δ = (const)
(2)
EI
or
This could not be predicted by a pure dimensional analysis. Strength-of-materials
­theory predicts that the value of the constant is 13.
An Alternate Step-by-Step Method by Ipsen [5]2
The pi theorem method, just explained and illustrated, is often called the repeating variable method of dimensional analysis. Select the repeating variables, add
one more, and you get a pi group. The writer likes it. This method is straightforward and systematically reveals all the desired pi groups. However, there are
drawbacks: (1) All pi groups contain the same repeating variables and might lack
variety or effectiveness, and (2) one must (sometimes laboriously) check that the
selected repeating variables do not form a pi group among themselves (see Prob.
P5.21).
Ipsen [5] suggests an entirely different procedure, a step-by-step method that
obtains all of the pi groups at once, without any counting or checking. One simply successively eliminates each dimension in the desired function by division or
multiplication. Let us illustrate with the same classical drag function proposed in
Eq. (5.1). Underneath the variables, write out the dimensions of each quantity.
F
{MLT −2 }
2
= fcn(L,
{L}
V,
{LT −1 }
ρ,
{ML−3 }
This method may be omitted without loss of continuity.
μ)
{ML−1T −1 }
(5.1)
310
Chapter 5 Dimensional Analysis and Similarity
There are three dimensions, {MLT}. Eliminate them successively by division
or multiplication by a variable. Start with mass {M}. Pick a variable that contains
mass and divide it into all the other variables with mass dimensions. We select
ρ, divide, and rewrite the function (5.1):
F
ρ
= fcn (L,
{L4T −2 }
{L}
μ
ρ)
{LT −1 } {L2T −1 }
V,
ρ
(5.1a)
We did not divide into L or V, which do not contain {M}. Equation (5.1a) at first
looks strange, but it contains five distinct variables and the same information as Eq.
(5.1).
We see that ρ is no longer important. Thus discard ρ, and now there are only
four variables. Next, eliminate time {T} by dividing the time-containing variables
by suitable powers of, say, V. The result is
F
= fcn (L,
ρV2
{L2}
{L}
μ
ρV )
{L}
V,
(5.1b)
Now we see that V is no longer relevant. Finally, eliminate {L} through division by, say, appropriate powers of L itself:
F
= fcn (L,
ρV2L2
μ
ρVL )
(5.1c)
{1}
{1}
Now L by itself is no longer relevant, and so discard it also. The result is
equivalent to Eq. (5.2):
μ
F
= fcn (
(5.2)
2 2
ρVL )
ρV L
In Ipsen’s step-by-step method, we find the force coefficient is a function solely
of the Reynolds number. We did no counting and did not find j. We just successively eliminated each primary dimension by division with the appropriate variables.
Recall Example 5.5, where we discovered, awkwardly, that the number of
repeating variables was less than the number of primary dimensions. Ipsen’s
method avoids this preliminary check. Recall the beam-deflection problem proposed in Example 5.5 and the various dimensions:
δ = f(P,
{L} {MLT−2 }
L,
{L}
I,
{L4 }
E)
{ML−1T−2 }
For the first step, let us eliminate {M} by dividing by E. We only have to divide
into P:
P
δ=f( ,
E
{L} {L2 }
L,
I, E)
{L} {L4 }
5.3 The Pi Theorem 311
We see that we may discard E as no longer relevant, and the dimension {T} has
vanished along with {M}. We need only eliminate {L} by dividing by, say, p­ owers
of L itself:
δ
P
I
= fcn ( 2 , L, 4 )
L
EL
L
{1}
{1}
{1}
Discard L itself as now irrelevant, and we obtain Answer (1) to Example 5.5:
δ
P I
= fcn ( 2 , 4 )
L
EL L
Ipsen’s approach is again successful. The fact that {M} and {T} vanished in the
same division is proof that there are only two repeating variables this time, not
the three that would be inferred by the presence of {M}, {L}, and {T}.
EXAMPLE 5.6
The leading-edge aerodynamic moment MLE on a supersonic airfoil is a function of its
chord length C, angle of attack α, and several air parameters: approach velocity V,
density ρ, speed of sound a, and specific-heat ratio k (Fig. E5.6). There is a very weak
effect of air ­viscosity, which is neglected here.
MLE
C
V
α
E5.6
Use Ipsen’s method to rewrite this function in dimensionless form.
Solution
Write out the given function and list the variables’ dimensions {MLT} underneath:
MLE = fcn(C,
{ML2/T2 }
{L}
α,
{1}
V,
{L/T}
ρ,
{M/L3 }
a,
{L/T}
k)
{1}
Two of them, α and k, are already dimensionless. Leave them alone; they will be pi
groups in the final function. You can eliminate any dimension. We choose mass {M}
and divide by ρ:
MLE
= fcn(C,
ρ
{L5/T2 } {L}
α,
V,
{1}
{L/T}
ρ,
a,
k)
{L/T} {1}
312
Chapter 5 Dimensional Analysis and Similarity
Recall Ipsen’s rules: Only divide into variables containing mass, in this case only MLE,
and then discard the divisor, ρ. Now eliminate time {T} by dividing by appropriate
powers of a:
V
= fcn (C, α,
, a,
a
ρa
{L3 }
{L} {1} {1}
k)
MLE
2
{1}
3
Finally, eliminate {L} on the left side by dividing by C :
MLE
ρa2C3
{1}
= fcn (C,
V
, k)
a
{1} {1} {1}
α,
We end up with four pi groups and recognize V/a as the Mach number, Ma. In aerodynamics, the dimensionless moment is often called the moment coefficient, CM. Thus
our final result could be written in the compact form
CM = fcn(α, Ma, k)
Ans.
Comments: Our analysis is fine, but experiment and theory and physical reasoning
all indicate that MLE varies more strongly with V than with a. Thus aerodynamicists
commonly define the moment coefficient as CM = MLE/(ρV2C3) or something similar.
We will study the analysis of supersonic forces and moments in Chap. 9.
5.4 Nondimensionalization of the Basic Equations
We could use the pi theorem method of the previous section to analyze problem
after problem after problem, finding the dimensionless parameters that govern in
each case. Textbooks on dimensional analysis [for example, 5] do this. An alternative and very powerful technique is to attack the basic equations of flow from
Chap. 4. Even though these equations cannot be solved in general, they will reveal
basic dimensionless parameters, such as the Reynolds number, in their proper
form and proper position, giving clues to when they are negligible. The boundary
conditions must also be nondimensionalized.
Let us briefly apply this technique to the incompressible flow continuity and
momentum equations with constant viscosity:
Continuity:
Navier–Stokes:
∇ · V = 0(5.9a)
ρ
dV
= ρg − ∇p + μ∇ 2V(5.9b)
dt
Typical boundary conditions for these two equations are (Sect. 4.6)
Fixed solid surface:
V = 0
Inlet or outlet:
Free surface, z = η: w =
Known V, p
dη
dt
−1
p = pa − Υ(R−1
x + Ry )
(5.10)
5.4 Nondimensionalization of the Basic Equations 313
We omit the energy equation (4.78) and assign its dimensionless form in the
problems (Prob. P5.43).
Equations (5.9) and (5.10) contain the three basic dimensions M, L, and T. All
variables p, V, x, y, z, and t can be nondimensionalized by using density and two
reference constants that might be characteristic of the particular fluid flow:
Reference velocity = U
Reference length = L
For example, U may be the inlet or upstream velocity and L the diameter of a
body immersed in the stream.
Now define all relevant dimensionless variables, denoting them by an asterisk:
V* =
x* =
x
L
V
U
y* =
t* =
tU
L
y
L
∇* = L∇
z
R
R* = L
L
p + ρgz
z* =
p* =
(5.11)
ρU2
All these are fairly obvious except for p*, where we have introduced the piezometric pressure, assuming that z is up. This is a hindsight idea suggested by
Bernoulli’s equation (3.54).
Since ρ, U, and L are all constants, the derivatives in Eqs. (5.9) can all be
handled in dimensionless form with dimensional coefficients. For example,
∂u ∂ (Uu*)
U ∂u*
=
=
∂x
∂ (Lx*)
L ∂x*
Substitute the variables from Eqs. (5.11) into Eqs. (5.09) and (5.10) and divide
through by the leading dimensional coefficient, U/L for the continuity equation,
and ρU2/L for the momentum equations. Here are the resulting dimensionless
equations of motion:
Continuity:
∇* · V* = 0 (5.12a)
Momentum:
μ
dV*
= −∇*p* +
∇*2 (V*) dt*
ρUL
(5.12b)
The dimensionless boundary conditions are:
Fixed solid surface:
Inlet or outlet:
Free surface, z* = η*:
V* = 0
Known V*, p*
dη*
dt*
pa
gL
Υ
p* =
+ 2 z* −
(R*x −1 + R*y −1 )
2
ρU
U
ρU2L
w* =
(5.13)
314
Chapter 5 Dimensional Analysis and Similarity
These equations reveal a total of four dimensionless parameters, one in the
Navier–Stokes equation and three in the free-surface-pressure boundary condition.
Dimensionless Parameters
In the continuity equation there are no parameters. The Navier–Stokes equation
contains one, generally accepted as the most important parameter in fluid mechanics:
Reynolds number Re =
ρUL
μ
It is named after Osborne Reynolds (1842–1912), a British engineer who first
proposed it in 1883 (Ref. 4 of Chap. 6). The Reynolds number is always important,
with or without a free surface, and can be neglected only in flow regions away
from high-velocity gradients—for example, away from solid surfaces, jets, or wakes.
The no-slip and inlet-exit boundary conditions contain no parameters. The
free-surface-pressure condition contains three:
pa
Euler number (pressure coefficient) Eu =
ρU2
This is named after Leonhard Euler (1707–1783) and is rarely important unless
the pressure drops low enough to cause vapor formation (cavitation) in a liquid.
The Euler number is often written in terms of pressure differences: Eu = Δp/(ρU2).
If Δp involves vapor pressure pυ, it is called the cavitation number Ca = (pa − pυ)/
(ρU2). Cavitation problems are surprisingly common in many water flows.
The second free-surface parameter is much more important:
Froude number Fr =
U2
gL
It is named after William Froude (1810–1879), a British naval architect who, with
his son Robert, developed the ship-model towing-tank concept and proposed similarity rules for free-surface flows (ship resistance, surface waves, open channels). The
Froude number is the dominant effect in free-surface flows. It can also be important
in stratified flows, where a strong density difference exists without a free surface.
For example, see Ref. [37]. Chapter 10 investigates Froude number effects in detail.
The final free-surface parameter is
Weber number We =
ρU2L
Υ
It is named after Moritz Weber (1871–1951) of the Polytechnic Institute of Berlin,
who developed the laws of similitude in their modern form. It was Weber who
named Re and Fr after Reynolds and Froude. The Weber number is important only
if it is of order unity or less, which typically occurs when the surface curvature is
comparable in size to the liquid depth, such as in droplets, capillary flows, ripple
waves, and very small hydraulic models. If We is large, its effect may be neglected.
If there is no free surface, Fr, Eu, and We drop out entirely, except for the possibility of cavitation of a liquid at very small Eu. Thus, in low-speed viscous flows with
no free surface, the Reynolds number is the only important dimensionless parameter.
5.4 Nondimensionalization of the Basic Equations 315
Compressibility Parameters
In high-speed flow of a gas there are significant changes in pressure, density, and
temperature that must be related by an equation of state such as the perfect-gas
law, Eq. (1.10). These thermodynamic changes introduce two additional dimensionless parameters mentioned briefly in earlier chapters:
Mach number Ma =
cp
U
Specific-heat ratio k =
a
cυ
The Mach number is named after Ernst Mach (1838–1916), an Austrian physicist. The effect of k is only slight to moderate, but Ma exerts a strong effect on
compressible flow properties if it is greater than about 0.3. These effects are
studied in Chap. 9.
Oscillating Flows
If the flow pattern is oscillating, a seventh parameter enters through the inlet
boundary condition. For example, suppose that the inlet stream is of the form
u = U cos ωt
Nondimensionalization of this relation results in
u
ωL
= u* = cos ( t*)
U
U
The argument of the cosine contains the new parameter
Strouhal number St =
ωL
U
The dimensionless forces and moments, friction, and heat transfer, and so on of
such an oscillating flow would be a function of both Reynolds and Strouhal
numbers. This parameter is named after V. Strouhal, a German physicist who
experimented in 1878 with wires singing in the wind.
Some flows that you might guess to be perfectly steady actually have an oscillatory pattern that is dependent on the Reynolds number. An example is the
periodic vortex shedding behind a blunt body immersed in a steady stream of
velocity U. Figure 5.1a shows an array of alternating vortices shed from a ­circular
cylinder immersed in a steady crossflow. This regular, periodic shedding is called
a Kármán vortex street, after T. von Kármán, who explained it theoretically in
1912. The shedding occurs in the range 102 < Re < 107, with an average ­Strouhal
number ωd/(2πU) ≈ 0.21. Figure 5.1b shows measured shedding frequencies.
Resonance can occur if a vortex shedding frequency is near a body’s structural
vibration frequency. Electric transmission wires sing in the wind, undersea mooring lines gallop at certain current speeds, and slender structures flutter at critical
wind or vehicle speeds. A striking example is the disastrous failure of the Tacoma
Narrows suspension bridge in 1940, when wind-excited vortex shedding caused
resonance with the natural torsional oscillations of the bridge. The problem was
magnified by the bridge deck nonlinear stiffness, which occurred when the
­hangers went slack during the oscillation.
316
Chapter 5 Dimensional Analysis and Similarity
(a)
0.4
Data spread
St = ω d
2π U
0.3
0.2
0.1
Fig. 5.1 Vortex shedding from a
­circular cylinder: (a) vortex street
behind a circular cylinder (Courtesy
of U.S. Navy); (b) experimental
shedding frequencies (data from
Refs. 24 and 25).
0
10
10 2
10 3
10 4
ρ Ud
Re =
µ
10 5
10 6
107
(b)
Other Dimensionless Parameters
We have discussed seven important parameters in fluid mechanics, and there are
others. Four additional parameters arise from nondimensionalization of the energy
equation (4.78) and its boundary conditions. These four (Prandtl number, Eckert
number, Grashof number, and wall temperature ratio) are listed in Table 5.1 just
in case you fail to solve Prob. P5.43. Another important and perhaps surprising
parameter is the wall roughness ratio ε/L (in Table 5.2).3 Slight changes in surface
3
Roughness is easy to overlook because it is a slight geometric effect that does not appear in
the equations of motion. It is a boundary condition that one might forget.
5.4 Nondimensionalization of the Basic Equations 317
Table 5.2 Dimensionless Groups
in Fluid Mechanics
Parameter
Definition
Re =
Mach number
Ma 5
Froude number
Fr 5
Weber number
We =
Rossby number
Ro 5
Cavitation number
(Euler number)
Ca =
Prandtl number
Pr =
Eckert number
Ec 5
Specific-heat ratio
k=
Strouhal number
St =
Roughness ratio
ε
L
Grashof number
Gr =
Rayleigh number
Ra =
Temperature ratio
Tw
T0
Pressure coefficient
Cp =
Lift coefficient
CL =
Drag coefficient
CD =
Friction factor
f=
Skin friction coefficient cf =
μ
Inertia
Viscosity
U
a
Sound speed
ρUL
Reynolds number
Qualitative ratio
of effects
Flow speed
U2
gL
ρU2L
Υ
U
V earth L
p − pυ
1
2
2 ρU
Inertia
Surface tension
Free-surface flow
Flow velocity
Coriolis effect
Pressure
Inertia
Dissipation
Conduction
U2
cpT0
Kinetic energy
Enthalpy
cp
Enthalpy
cυ
Internal energy
Oscillation
Mean speed
Wall roughness
Body length
Buoyancy
2
Viscosity
βΔTgL3ρ2cp
Buoyancy
μk
Viscosity
μ
Wall temperature
Stream temperature
p − p∞
Static pressure
1
2
2 ρU
Dynamic pressure
L
1
2
2 ρU A
Lift force
Dynamic force
D
Drag force
1
2
2 ρU A
Dynamic force
hf
2
(V /2g) (L/d)
τwall
ρV 2/2
Compressible flow
Free-surface flow
k
βΔTgL3ρ2
Almost always
Inertia
Gravity
μcp
ωL
U
Importance
Geophysical flows
Cavitation
Heat convection
Dissipation
Compressible flow
Oscillating flow
Turbulent, rough walls
Natural convection
Natural convection
Heat transfer
Aerodynamics, hydrodynamics
Aerodynamics, hydrodynamics
Aerodynamics, hydrodynamics
Friction head loss
Velocity head
Pipe flow
Wall shear stress
Dynamic pressure
Boundary layer flow
318
Chapter 5 Dimensional Analysis and Similarity
5
4
Cylinder
length effect
Transition to turbulent
boundary layer
3
(10 4 < Re < 10 5)
CD
2
Cylinder (two-dimensional)
1
L/d
CD
∞
40
20
10
5
3
2
1
1.20
0.98
0.91
0.82
0.74
0.72
0.68
0.64
Sphere
0
10
10 2
10 3
10 4
ρ Ud
Red =
µ
(a)
10 5
10 6
10 7
1.5
1.0
Fig. 5.2 The proof of practical
­dimensional analysis: drag
­coefficients of a cylinder and
sphere: (a) drag coefficient of a
smooth cylinder and sphere (data
from many sources); (b) increased
roughness causes earlier transition
to a turbulent boundary layer.
CD 0.7
0.5
0.3
10 4
Cylinder:
 = 0.02
d
0.009
0.007
0.004
0.002
0.0005
_L = ∞
d
Smooth
10 5
Red
10 6
(b)
roughness have a striking effect in the turbulent flow or high-Reynolds-number
range, as we shall see in Chap. 6 and in Fig. 5.2.
This book is primarily concerned with Reynolds-, Mach-, and Froude-number
effects, which dominate most flows. Note that we discovered these parameters
(except ε/L) simply by nondimensionalizing the basic equations without actually
solving them.
If the reader is not satiated with the 19 parameters given in Table 5.2, Ref. 26
contains a list of over 1200 dimensionless parameters in use in engineering and
science.
A Successful Application
Dimensional analysis is fun, but does it work? Yes, if all important variables are
included in the proposed function, the dimensionless function found by
5.4 Nondimensionalization of the Basic Equations 319
d­ imensional analysis will collapse all the data onto a single curve or set of
curves.
An example of the success of dimensional analysis is given in Fig. 5.2 for the
measured drag on smooth cylinders and spheres. The flow is normal to the axis
of the cylinder, which is extremely long, L/d → ∞. The data are from many
sources, for both liquids and gases, and include bodies from several meters in
diameter down to fine wires and balls less than 1 mm in size. Both curves in Fig.
5.2a are entirely experimental; the analysis of immersed body drag is one of the
weakest areas of modern fluid mechanics theory. Except for digital computer
calculations, there is little theory for cylinder and sphere drag except creeping
flow, Re < 1.
The concept of a fluid-caused drag force on bodies is covered extensively in
Chap. 7. Drag is the fluid force parallel to the oncoming stream—see Fig. 7.10
for details.
The Reynolds number of both bodies is based on diameter, hence the notation
Red. But the drag coefficients are defined differently:
drag
CD = µ
1
2
1
2
ρU 2Ld
drag
ρU 2 14πd 2
cylinder
(5.14)
sphere
They both have a factor 12 because the term 12ρU 2 occurs in Bernoulli’s equation,
and both are based on the projected area—that is, the area one sees when looking
toward the body from upstream. The usual definition of CD is thus
CD =
drag
1
2
2
ρU (projected area)
(5.15)
However, one should carefully check the definitions of CD, Re, and the like before
using data in the literature. Airfoils, for example, use the planform area.
Figure 5.2a is for long, smooth cylinders. If wall roughness and cylinder length
are included as variables, we obtain from dimensional analysis a complex threeparameter function:
ε L
CD = f (Red, , )
d d
(5.16)
To describe this function completely would require 1000 or more experiments or
CFD results. Therefore it is customary to explore the length and roughness effects
separately to establish trends.
The table with Fig. 5.2a shows the length effect with zero wall roughness. As
length decreases, the drag decreases by up to 50 percent. Physically, the pressure
is “relieved” at the ends as the flow is allowed to skirt around the tips instead of
deflecting over and under the body.
Figure 5.2b shows the effect of wall roughness for an infinitely long cylinder.
The sharp drop in drag occurs at lower Red as roughness causes an earlier transition to a turbulent boundary layer on the surface of the body. Roughness has the
320
Chapter 5 Dimensional Analysis and Similarity
same effect on sphere drag, a fact that is exploited in sports by deliberate dimpling
of golf balls to give them less drag at their flight Red ≈ 105. See Fig. D5.2.
Figure 5.2 is a typical experimental study of a fluid mechanics problem, aided
by dimensional analysis. As time and money and demand allow, the complete
three-parameter relation (5.16) could be filled out by further experiments.
EXAMPLE 5.7
A smooth cylinder, 1 cm in diameter and 20 cm long, is tested in a wind tunnel for a
crossflow of 45 m/s of air at 20°C and 1 atm. The measured drag is 2.2 ± 0.1 N. (a)
Does this data point agree with the data in Fig. 5.2? (b) Can this data point be used to
predict the drag of a chimney 1 m in diameter and 20 m high in winds at 20°C and
1 atm? If so, what is the recommended range of wind velocities and drag forces for
this data point? (c) Why are the answers to part (b) always the same, regardless of the
chimney height, as long as L = 20d?
Solution
(a) For air at 20°C and 1 atm, take ρ = 1.2 kg/m3 and µ = 1.8 E−5 kg/(m-s). Since the test
cylinder is short, L /d = 20, it should be compared with the tabulated value CD ≈ 0.91 in
the table to the right of Fig. 5.2a. First calculate the Reynolds number of the test cylinder:
Red =
ρUd (1.2 kg/m3 ) (45 m/s) (0.01 m)
=
= 30,000
μ
1.8E−5 kg/(m − s)
Yes, this is in the range 104 < Re < 105 listed in the table. Now calculate the test drag
coefficient:
F
2.2 N
= 0.905
=
2
3
(1/2)ρU Ld (1/2) (1.2 kg/m ) (45 m/s) 2 (0.2 m) (0.01 m)
CD,test =
Yes, this is close, and certainly within the range of ±5 percent stated by the test results.
Ans. (a)
(b) Since the chimney has L/d = 20, we can use the data if the Reynolds number range
is correct:
104 <
(1.2 kg/m3 )Uchimney (1 m)
1.8 E−5 kg/(m · s)
< 105 if 0.15
m
m
< Uchimney < 1.5
s
s
These are negligible winds, so the test data point is not very useful
The drag forces in this range are also negligibly small:
Ans. (b)
Fmin = CD
1.2 kg/m3
ρ 2
2
Umin Ld = (0.91) (
) (0.15 m/s) (20 m) (1 m) = 0.25 N
2
2
Fmax = CD
1.2 kg/m3
ρ 2
2
Umax Ld = (0.91) (
) (1.5 m/s) (20 m) (1 m) = 25 N
2
2
(c) Try this yourself. Choose any 20:1 size for the chimney, even something silly like
20 mm:1 mm. You will get the same results for U and F as in part (b) above. This is
5.5 Modeling and Similarity 321
because the product Ud occurs in Red and, if L = 20d, the same product occurs in the
drag force. For example, for Re = 104,
ρ
ρ
ρ
ρ 104μ 2
μ
Ud = 104 then F = CD U2Ld = CD U2 (20d )d = 20CD (Ud ) 2 = 20CD (
ρ
ρ )
2
2
2
2
The answer is always Fmin = 0.25 N. This is an algebraic quirk that seldom occurs.
EXAMPLE 5.8
Telephone wires are said to “sing” in the wind. Consider a wire of diameter 8 mm. At
what sea-level wind velocity, if any, will the wire sing a middle C note?
Solution
For sea-level air take ν ≈ 1.5 E−5 m2/s. For nonmusical readers, middle C is 262 Hz.
Measured shedding rates are plotted in Fig. 5.1b. Over a wide range, the Strouhal number
is ­approximately 0.2, which we can take as a first guess. Note that (ω/2π) = f, the shedding frequency. Thus
St =
fd (262 s−1 ) (0.008 m)
=
≈ 0.2
U
U
m
U ≈ 10.5
s
Now check the Reynolds number to see if we fall into the appropriate range:
Red =
Ud (10.5 m/s) (0.008 m)
=
≈ 5600
ν
1.5 E−5 m2/s
In Fig. 5.1b, at Re = 5600, maybe St is a little higher, at about 0.21. Thus a slightly
improved estimate is
Uwind = (262) (0.008)/(0.21) ≈ 10.0 m/s
5.5 Modeling and Similarity
Ans.
So far we have learned about dimensional homogeneity and the pi theorem
method, using power products, for converting a homogeneous physical relation to
dimensionless form. This is straightforward mathematically, but certain engineering difficulties need to be discussed.
First, we have more or less taken for granted that the variables that affect the
process can be listed and analyzed. Actually, selection of the important variables
requires considerable judgment and experience. The engineer must decide, for
example, whether viscosity can be neglected. Are there significant temperature
effects? Is surface tension important? What about wall roughness? Each pi group
that is retained increases the expense and effort required. Judgment in selecting
variables will come through practice and maturity; this book should provide some
of the necessary experience.
322
Chapter 5 Dimensional Analysis and Similarity
Once the variables are selected and the dimensional analysis is performed, the
experimenter seeks to achieve similarity between the model tested and the prototype to be designed. With sufficient testing, the model data will reveal the desired
dimensionless function between variables:
Π 1 = f(Π 2, Π 3, … Π k )
(5.17)
With Eq. (5.17) available in chart, graphical, or analytical form, we are in a position to ensure complete similarity between model and prototype. A formal statement would be as follows:
Flow conditions for a model test are completely similar if all relevant dimensionless parameters have the same corresponding values for the model and the
prototype.
This follows mathematically from Eq. (5.17). If Π2m = Π2p, Π3m = Π3p, and so
forth, Eq. (5.17) guarantees that the desired output Π1m will equal Π1p. But this
is easier said than done, as we now discuss. There are specialized texts on model
testing [27–29].
Instead of complete similarity, the engineering literature speaks of particular
types of similarity, the most common being geometric, kinematic, dynamic, and
thermal. Let us consider each separately.
Geometric Similarity
Geometric similarity concerns the length dimension {L} and must be ensured
before any sensible model testing can proceed. A formal definition is as
­follows:
A model and prototype are geometrically similar if and only if all body dimensions in all three coordinates have the same linear scale ratio.
Note that all length scales must be the same. It is as if you took a photograph
of the prototype and reduced it or enlarged it until it fitted the size of the
model. If the model is to be made one-tenth the prototype size, its length,
width, and height must each be one-tenth as large. Not only that, but also its
entire shape must be one-tenth as large, and technically we speak of homologous points, which are points that have the same relative location. For example,
the nose of the prototype is homologous to the nose of the model. The left
wingtip of the prototype is homologous to the left wingtip of the model. Then
geometric similarity requires that all homologous points be related by the same
linear scale ratio. This applies to the fluid geometry as well as the model
geometry.
All angles are preserved in geometric similarity. All flow directions are preserved. The orientations of model and prototype with respect to the surroundings must be identical.
Figure 5.3 illustrates a prototype wing and a one-tenth-scale model. The model
lengths are all one-tenth as large, but its angle of attack with respect to the free
stream is the same for both model and prototype: 10° not 1°. All physical details
5.5 Modeling and Similarity 323
*
40 m
1m
Homologous
points
a
a
4m
10°
0.1 m
10°
Vp
8m
Fig. 5.3 Geometric similarity in
model testing: (a) prototype;
(b) one-tenth-scale model.
Vm
*
0.8 m
(b)
(a)
on the model must be scaled, and some are rather subtle and sometimes overlooked:
1. The model nose radius must be one-tenth as large.
2. The model surface roughness must be one-tenth as large.
3. If the prototype has a 5-mm boundary layer trip wire 1.5 m from the leading
edge, the model should have a 0.5-mm trip wire 0.15 m from its leading edge.
4. If the prototype is constructed with protruding fasteners, the model should
have homologous protruding fasteners one-tenth as large.
And so on. Any departure from these details is a violation of geometric similarity and must be justified by experimental comparison to show that the prototype
behavior was not significantly affected by the discrepancy.
Models that appear similar in shape but that clearly violate geometric similarity should not be compared except at your own risk. Figure 5.4 illustrates this
point. The spheres in Fig. 5.4a are all geometrically similar and can be tested
with a high expectation of success if the Reynolds number, Froude number, or
the like is matched. But the ellipsoids in Fig. 5.4b merely look similar. They
V2
V1
Huge
sphere
V3
Large
sphere
V4
Medium
sphere
Tiny
sphere
(a)
V1
Fig. 5.4 Geometric similarity and
dissimilarity of flows: (a) similar;
(b) dissimilar.
V2
Large 4:1
ellipsoid
V3
Medium 3.5:1
ellipsoid
(b)
Small 3:1
ellipsoid
324
Chapter 5 Dimensional Analysis and Similarity
actually have different linear scale ratios and therefore cannot be compared in a
rational manner, even though they may have identical Reynolds and Froude numbers and so on. The data will not be the same for these ellipsoids, and any attempt
to “compare” them is a matter of rough engineering judgment.
Kinematic Similarity
Kinematic similarity requires that the model and prototype have the same length
scale ratio and the same time scale ratio. The result is that the velocity scale ratio
will be the same for both. As Langhaar [4] states it:
The motions of two systems are kinematically similar if homologous particles
lie at homologous points at homologous times.
Length scale equivalence simply implies geometric similarity, but time scale equivalence may require additional dynamic considerations such as equivalence of the
Reynolds and Mach numbers.
One special case is incompressible frictionless flow with no free surface, as
sketched in Fig. 5.5a. These perfect-fluid flows are kinematically similar with
independent length and time scales, and no additional parameters are necessary (see
Chap. 8 for further details).
Froude Scaling
Frictionless flows with a free surface, as in Fig. 5.5b, are kinematically similar
if their Froude numbers are equal:
Vp2
Vm2
Frm =
=
= Frp
gLm gLp
(5.18)
Note that the Froude number contains only length and time dimensions and hence
is a purely kinematic parameter that fixes the relation between length and time.
From Eq. (5.18), if the length scale is
Lm = αLp
(5.19)
where α is a dimensionless ratio, the velocity scale is
Vm
Lm 1∕2
=( ) =
Vp
Lp
√α
(5.20)
and the time scale is
Tm Lm ∕Vm
=
=
Tp
Lp ∕Vp
√α
(5.21)
These Froude-scaling kinematic relations are illustrated in Fig. 5.5b for wave
motion modeling. If the waves are related by the length scale α, then the wave
period, propagation speed, and particle velocities are related by √α.
If viscosity, surface tension, or compressibility is important, kinematic similarity depends on the achievement of dynamic similarity.
5.5 Modeling and Similarity 325
V1p
V1m = βV1p
Dp
V∞ p
V∞ m = βV∞p
Dm =
α Dp
Model
V2 m = βV2 p
V2p
Prototype
(a)
λp
Prototype
waves:
Cp
Hp
Period Tp
Vp
λ m = α λp
Hm = α Hp
Cm = Cp α
Model
waves:
Period Tm = Tpα
Vm = Vp α
(b)
Fig. 5.5 Frictionless low-speed flows are kinematically similar: (a) Flows with no free surface are kinematically similar with independent length and time scale ratios; (b) free-surface
flows are kinematically similar with length and time scales related by the Froude number.
Dynamic Similarity
Dynamic similarity exists when the model and the prototype have the same length
scale ratio, time scale ratio, and force scale (or mass scale) ratio. Again geometric similarity is a first requirement; without it, proceed no further. Then dynamic
similarity exists, simultaneous with kinematic similarity, if the model and prototype force and pressure coefficients are identical. This is ensured if
1. For compressible flow, the model and prototype Reynolds number and Mach
number and specific-heat ratio are correspondingly equal.
2. For incompressible flow
a. With no free surface: model and prototype Reynolds numbers are equal.
b. With a free surface: model and prototype Reynolds number, Froude
­number, and (if necessary) Weber number and cavitation number are
­correspondingly equal.
326
Chapter 5 Dimensional Analysis and Similarity
Fpp
Fgp
Fip
Ffp
Fig. 5.6 Dynamic similarity in
sluice gate flow. Model and
­prototype yield identical homologous force polygons if the Reynolds
and Froude numbers are the same
corresponding values: (a) prototype;
(b) model.
Fpm
Fim
a
Fgm
Ffm
a'
(a)
(b)
Mathematically, Newton’s law for any fluid particle requires that the sum of the
pressure force, gravity force, and friction force equal the acceleration term, or
inertia force,
Fp + Fg + Ff = Fi
The dynamic similarity laws listed above ensure that each of these forces will be
in the same ratio and have equivalent directions between model and prototype.
Figure 5.6 shows an example for flow through a sluice gate. The force polygons
at homologous points have exactly the same shape if the Reynolds and Froude
numbers are equal (neglecting surface tension and cavitation, of course). ­Kinematic
similarity is also ensured by these model laws.
Incomplete Similarity
The perfect dynamic similarity shown in Fig. 5.6 is more of a dream than a reality because true equivalence of Reynolds and Froude numbers can be achieved
only by dramatic changes in fluid properties, whereas in fact most model testing
is simply done with water or air, the cheapest fluids available.
First consider hydraulic model testing with a free surface. Dynamic similarity
requires equivalent Froude numbers, Eq. (5.18), and equivalent Reynolds numbers:
Vm Lm Vp L p
=
νm
νp
(5.22)
But both velocity and length are constrained by the Froude number, Eqs. (5.19)
and (5.20). Therefore, for a given length scale ratio α, Eq. (5.22) is true only if
νm Lm Vm
=
= α √α = α3/2
νp
Lp Vp
(5.23)
For example, for a one-tenth-scale model, α = 0.1 and α3/2 = 0.032. Since νp is
undoubtedly water, we need a fluid with only 0.032 times the kinematic viscosity
5.5 Modeling and Similarity 327
Range
of Rem
log CD
Fig. 5.7 Reynolds-number extrapolation, or scaling, of hydraulic data
with equal Froude numbers.
Range
of Re p
Power-law
extrapolation
Uncertainty
in prototype
data estimate
Model
data:
105
106
log Re
107
108
of water to achieve dynamic similarity. Referring to Table 1.4, we see that this is
impossible: Even mercury has only one-ninth the kinematic viscosity of water, and
a mercury hydraulic model would be expensive and bad for your health. In practice,
water is used for both the model and the prototype, and the Reynolds number similarity (5.22) is unavoidably violated. The Froude number is held constant since it
is the dominant parameter in free-surface flows. Typically the Reynolds number of
the model flow is too small by a factor of 10 to 1000. As shown in Fig. 5.7, the
low-Reynolds-number model data are used to estimate by extrapolation the desired
high-Reynolds-number prototype data. As the figure indicates, there is obviously
considerable uncertainty in using such an extrapolation, but there is no other practical alternative in hydraulic model testing.
Second, consider aerodynamic model testing in air with no free surface. The
important parameters are the Reynolds number and the Mach number. Equation
(5.22) should be satisfied, plus the compressibility criterion
Vm Vp
= am ap
(5.24)
Elimination of Vm/Vp between (5.22) and (5.24) gives
νm Lm am
=
νp
Lp ap
(5.25)
Since the prototype is no doubt an air operation, we need a wind-tunnel fluid of
low viscosity and high speed of sound. Hydrogen is the only practical example,
but clearly it is too expensive and dangerous. Therefore, wind tunnels normally
operate with air as the working fluid. Cooling and pressurizing the air will bring
Eq. (5.25) into better agreement but not enough to satisfy a length scale reduction
of, say, one-tenth. Therefore Reynolds number scaling is also commonly violated
in aerodynamic testing, and an extrapolation like that in Fig. 5.7 is required here
also.
There are specialized monographs devoted entirely to wind tunnel testing: low
speed [33], high speed [34], and a detailed general discussion [35]. The following
example illustrates modeling discrepancies in aeronautical testing.
328
Chapter 5 Dimensional Analysis and Similarity
EXAMPLE 5.9
A prototype airplane, with a chord length of 1.6 m, is to fly at Ma = 2 at 10 km standard altitude. A one-eighth scale model is to be tested in a helium wind tunnel at 100°C
and 1 atm. Find the helium test section velocity that will match (a) the Mach number
or (b) the Reynolds number of the prototype. In each case criticize the lack of dynamic
similarity. (c) What high pressure in the helium tunnel will match both the Mach and
Reynolds numbers? (d) Why does part (c) still not achieve dynamic similarity?
Solution
For helium, from Table A.4, R = 2077 m2/(s2-K), k = 1.66, and estimate µHe ≈ 2.32
E−5 kg/(m · s) from the power-law, n = 0.67, in the table. (a) Calculate the helium speed
of sound and velocity:
aHe =
= √ (1.66) (2077 m2/s2K) × (373 K) = 1134 m/s
VHe
VHe
Maair = MaHe = 2.0 =
=
aHe 1134 m/s
m
VHe = 2268 Ans. (a)
s
√ (kRT) He
For dynamic similarity, the Reynolds numbers should also be equal. From Table A.6
at an altitude of 10,000 m, read ρair = 0.4125 kg/m3, aair = 299.5 m/s, and estimate µair
≈ 1.48 E−5 kg/m · s from the power-law, n = 0.7, in Table A.4. The air velocity is
Vair = (Ma)(aair) = 2(299.5) = 599 m/s. The model chord length is (1.6 m)/8 = 0.2 m.
The helium density is ρHe = (p/RT)He = (101,350 Pa)/[(2077 m2/s2 K)(373 K)] =
0.131 kg/m3. Now calculate the two Reynolds numbers:
ReC,air =
ReC,He =
(0.4125 kg/m3 ) (599 m/s) (1.6 m)
ρVC
= 26.6 E6
` =
μ air
1.48 E−5 kg/(m · s)
(0.131 kg/m3 ) (2268 m/s) (0.2 m)
ρVC
` =
= 2.56 E6
μ He
2.32 E−5 kg/(m · s)
The model Reynolds number is 10 times less than the prototype. This is typical when
using small-scale models. The test results must be extrapolated for Reynolds number
effects.
(b) Now ignore Mach number and let the model Reynolds number match the
prototype:
(0.131 kg/m3 )VHe (0.2 m)
2.32 E−5 kg/(ms )
m
VHe = 23,600 s
ReHe = Reair = 26.6 E6 =
Ans. (b)
This is ridiculous: a hypersonic Mach number of 21, suitable for escaping from the
earth’s gravity. One should match the Mach numbers and correct for a lower Reynolds
number.
(c) Match both Reynolds and Mach numbers by increasing the helium density:
Ma matches if
VHe = 2268
m
s
5.5 Modeling and Similarity 329
Then
ReHe = 26.6 E6 =
ρHe (2268 m/s) (0.2 m)
2.32 E−5 kg/(m · s)
Solve for
ρHe = 1.36
kg
m3
pHe = ρRT∣He = (1.36) (2077) (373) = 1.05 E6 Pa
Ans. (c)
A match is possible if we increase the tunnel pressure by a factor of ten, a daunting
task.
(d) Even with Ma and Re matched, we are still not dynamically similar because the
two gases have different specific-heat ratios: kHe = 1.66 and kair = 1.40. This discrepancy will cause substantial differences in pressure, density, and temperature throughout
supersonic flow.
Figure 5.8 shows the final hydraulic model design for the U.S. Army Corps
of Engineers Sacramento District’s Isabella Lake Dam Safety Modification Project
at Utah State University’s Water Research Laboratory in Logan, Utah. The 1:45
scale model is an essential part of the Corps’ pre-construction engineering and
Fig. 5.8 Hydraulic model of the Isabella Lake Dam Safety Modification Project. The model scale is 1:45, and was built in 2014 at
Utah State University’s Water Research Laboratory. (Courtesy of the U.S. Army photo by John Prettyman/Released.)
330
Chapter 5 Dimensional Analysis and Similarity
design phase and allows engineers to test the design against extreme storms and
improve it before construction begins.
For hydraulic models of larger scale, such as harbors, estuaries, and embayments, geometric similarity may be violated of necessity. The vertical scale will
be distorted to avoid Weber number effects. For example, the horizontal scale
may be 1:1000, while the vertical scale is only 1:100. Thus the model channel
may be deeper relative to its horizontal dimensions. Since deeper passages flow
more efficiently, the model channel bottom may be deliberately roughened to
create the friction level expected in the prototype.
EXAMPLE 5.10
The pressure drop due to friction for flow in a long, smooth pipe is a function of average flow velocity, density, viscosity, and pipe length and diameter: Δp = fcn(V, ρ, µ,
L, D). We wish to know how Δp varies with V. (a) Use the pi theorem to rewrite this
function in dimensionless form. (b) Then plot this function, using the following data
for three pipes and three fluids:
Q, m3/h
V, m/s*
0.3 4,680 680†
2.92 E–4†
1.06
0.6
22,300 680†
2.92 E–4†
2.12
9.0
1.0
70,800 680†
2.92 E–4†
3.54
2.0
4.0
1.0 2,080 998‡
0.0010‡
0.88
2.0
6.0
2.0
10,500 998‡
0.0010‡
1.77
2.0
8.0
3.1
30,400 998‡
0.0010‡
2.74
3.0
3.0
0.5 540
13,550§
1.56 E–3§
0.20
3.0
4.0
1.0 2,480
13,550§
1.56 E–3§
0.39
3.0
5.0
1.7 9,600
13,550§
1.56 E–3§
0.67
L, m
1.0
5.0
1.0
7.0
1.0
Δp, Pa
ρ, kg/m3
µ, kg/(m · s)
D, cm
*V = Q/A, A = πD2/4.
†Gasoline.
‡Water.
§Mercury.
(c) Suppose it is further known that Δp is proportional to L (which is quite true for long
pipes with well-rounded entrances). Use this information to simplify and improve the pi
theorem formulation. Plot the dimensionless data in this improved manner and comment
on the results.
Solution
There are six variables with three primary dimensions involved {MLT}. Therefore, we
expect that j = 6 − 3 = 3 pi groups. We are correct, for we can find three variables that
do not form a pi product (e.g., ρ, V, L). Carefully select three ( j) repeating variables,
but not including Δp or V, which we plan to plot versus each other. We select (ρ, µ,
5.5 Modeling and Similarity 331
D), and the pi theorem guarantees that three independent power-product groups will
occur:
Π 1 = ρaμbDc Δp
or
Π1 =
Π 2 = ρdμeDfV
2
ρD Δp
μ
Π2 =
2
Π 3 = ρgμhDiL
ρVD
μ
Π3 =
L
D
We have omitted the algebra of finding (a, b, c, d, e, f, g, h, i) by setting all exponents
to zero M0, L0, T0. Therefore, we wish to plot the dimensionless relation
ρD2 Δp
μ
2
= fcn (
ρVD L
,
μ D)
Ans. (a)
We plot Π1 versus Π2 with Π3 as a parameter. There will be nine data points. For
example, the first row in the data here yields
ρD2 Δp
μ
2
=
(680) (0.01) 2 (4680)
(2.92 E− 4) 2
= 3.73 E9
ρVD (680) (1.06) (0.01)
=
= 24,700
μ
2.92 E−4
L
= 500
D
The nine data points are plotted as the open circles in Fig. 5.9. The values of L/D
are listed for each point, and we see a significant length effect. In fact, if we connect
the only two points that have the same L/D (= 200), we could see (and cross-plot to
verify) that Δp increases linearly with L, as stated in the last part of the problem. Since
L occurs only in Π3 = L/D, the function Π1 = fcn(Π2, Π3) must reduce to Π1 = (L/D)
fcn(Π2), or simply a function involving only two parameters:
ρD3 Δp
Lμ2
= fcn (
ρVD
μ )
flow in a long pipe
Ans. (c)
We now modify each data point in Fig. 5.9 by dividing it by its L/D value. For example,
for the first row of data, ρD3 Δp/(Lµ2) = (3.73 E9)/500 = 7.46 E6. We replot these
900
700
L = 200
D
Fig. 5.9 Two different correlations of
the data in Example 5.10: Open circles
when plotting ρD2 Δp/µ2 versus ReD,
L/D is a parameter; once it is known
that Δp is proportional to L, a replot
(solid circles) of ρD3 Δp/(Lµ2) versus
ReD collapses into a single powerlaw curve.
500
300
400
200
1010 Π1
133
100
109
108
Π1 107
Π3
10 6
10 4
0.155 ReD1.75
ReD
1011
105
332
Chapter 5 Dimensional Analysis and Similarity
new data points as solid circles in Fig. 5.9. They correlate almost perfectly into a
straight-line power-law function:
ρD3 Δp
2
Lμ
≈ 0.155 (
ρVD 1.75
μ )
Ans. (c)
All newtonian smooth pipe flows should correlate in this manner. This example is a
variation of the first completely successful dimensional analysis, pipe-flow friction,
performed by Prandtl’s student Paul Blasius, who published a related plot in 1911. For
this range of (turbulent flow) Reynolds numbers, the pressure drop increases approximately as V1.75.
EXAMPLE 5.11
The smooth sphere data plotted in Fig. 5.2a represent dimensionless drag versus dimensionless viscosity, since (ρ, V, d) were selected as scaling or repeating variables. (a)
Replot these data to display the effect of dimensionless velocity on the drag. (b) Use
your new figure to predict the terminal (zero-acceleration) velocity of a 1-cm-diameter
steel ball (SG = 7.86) falling through water at 20°C.
Solution
∙ Assumptions: Fig 5.2a is valid for any smooth sphere in that Reynolds number range.
∙Approach (a): Form pi groups from the function F = fcn(d, V, ρ, µ) in such a way
that F is plotted versus V. The answer was already given in the discussion of Example
5.2, but let us review the steps. The proper scaling variables are (ρ, µ, d ), which do
not form a pi. Therefore j = 3, and we expect n − j = 5 − 3 = 2 pi groups. Skipping
the algebra, they arise as follows:
Π 1 = ρaμbd c F =
ρF
μ2
Π 2 = ρaμbd c V =
ρVd
μ
Ans. (a)
We may replot the data of Fig. 5.2a in this new form, noting that Π1 ≡ (π/8)(CD)
(Re)2. This replot is shown as Fig. 5.10. The drag increases rapidly with velocity up
to transition, where there is a slight drop, after which it increases more than ever. If
force is known, we may predict velocity from the figure, and vice versa.
∙ Property values for part (b): ρwater = 998 kg/m3 µwater = 0.001 kg/(m-s)
ρsteel = 7.86ρwater = 7844 kg/m3.
∙ Solution to part (b):
the sphere in water:
For terminal velocity, the drag force equals the net weight of
F = Wnet = (ρs − ρw )g
π 3
π
d = (7840 − 998) (9.81) ( ) (0.01) 3 = 0.0351 N
6
6
Therefore, the ordinate of Fig. 5.10 is known:
Falling steel sphere:
ρF
μ
2
=
(998 kg/m3 ) (0.0351 N)
≈ 3.5 E7
[0.001 kg/(m · s) ] 2
Summary 333
1011
1010
Transition:
109
108
ρF
µ2
=
π
C Re2
8 D
107
106
10 5
10 4
103
102
10
Fig. 5.10 Cross-plot of sphere drag
data from Fig. 5.2a to show dimensionless force versus dimensionless
velocity.
1
0.1
1
10 2
10 3
ρ Vd
Re =
µ
10
10 4
10 5
10 6
From Fig. 5.10, at ρF/µ2 ≈ 3.5 E7, a magnifying glass reveals that Red ≈ 2 E4. Then
a crude estimate of the terminal fall velocity is
ρVd
≈ 20,000
μ
or
V≈
20,000[0.001 kg/(m · s) ]
3
(998 kg/m ) (0.01 m)
≈ 2.0
m
s
Ans. (b)
∙ Comments: Better accuracy could be obtained by expanding the scale of Fig. 5.10
in the region of the given force coefficient. However, there is considerable uncertainty
in published drag data for spheres, so the predicted fall velocity is probably uncertain
by at least ±10 percent.
Note that we found the answer directly from Fig. 5.10. We could use Fig. 5.2a
also but would have to iterate between the ordinate and abscissa to obtain the final
result, since V is contained in both plotted variables.
Summary
Chapters 3 and 4 presented integral and differential methods of mathematical
analysis of fluid flow. This chapter introduces the third and final method: experimentation, as supplemented by the technique of dimensional analysis. Tests and
experiments are used both to strengthen existing theories and to provide useful
engineering results when theory is inadequate.
334
Chapter 5 Dimensional Analysis and Similarity
The chapter begins with a discussion of some familiar physical relations and how
they can be recast in dimensionless form because they satisfy the principle of dimensional homogeneity. A general technique, the pi theorem, is then presented for systematically finding a set of dimensionless parameters by grouping a list of variables
that govern any particular physical process. A second technique, Ipsen’s method, is
also described. Alternately, direct application of dimensional analysis to the basic
equations of fluid mechanics yields the fundamental parameters governing flow patterns: Reynolds number, Froude number, Prandtl number, Mach number, and others.
It is shown that model testing in air and water often leads to scaling difficulties for which compromises must be made. Many model tests do not achieve true
dynamic similarity. The chapter ends by pointing out that classic dimensionless
charts and data can be manipulated and recast to provide direct solutions to
problems that would otherwise be quite cumbersome and laboriously iterative.
Problems
Most of the problems herein are fairly straightforward. More
difficult or open-ended assignments are labeled with an asterisk. Problems labeled with a computer icon
may require
the use of a computer. The standard end-of-chapter problems
P5.1 to P5.91 (categorized in the problem list here) are followed by word problems W5.1 to W5.9, fundamentals of engineering exam problems FE5.1 to FE5.12, comprehensive
applied problems C5.1 to C5.5, and design projects D5.1 and
D5.2.
Problem Distribution
Section
Topic
Problems
5.1
5.2
5.3
5.4
5.4
5.5
5.5
5.5
Introduction
The principle of dimensional homogeneity
The pi theorem; Ipsen’s method
Nondimensionalizing the basic equations
Data for spheres, cylinders, other bodies
Scaling of model data
Froude and Mach number scaling
Inventive rescaling of the data
P5.1–P5.9
P5.10–P5.13
P5.14–P5.42
P5.43–P5.47
P5.48–P5.59
P5.60–P5.74
P5.75–P5.84
P5.85–P5.91
P5.3
P5.4
P5.5
Introduction; dynamic similarity
P5.1
P5.2
For axial flow through a circular tube, the Reynolds
number for transition to turbulence is approximately
2300 [see Eq. (6.2)], based on the diameter and average velocity. If d = 5 cm and the fluid is kerosene at
20°C, find the volume flow rate in m3/h that causes
transition.
A prototype automobile is designed for cold weather
in Denver, CO (−10°C, 83 kPa). Its drag force is to
be tested on a one-seventh-scale model in a wind tun-
P5.6
P5.7
nel at 150 mi/h, 20°C, and 1 atm. If the model and prototype are to satisfy dynamic similarity, what prototype
­velocity, in mi/h, needs to be matched? Comment on your
result.
The transfer of energy by viscous dissipation is dependent
upon viscosity µ, thermal conductivity k, stream velocity U,
and stream temperature T0. Group these quantities, if
­possible, into the dimensionless Brinkman number, which
is proportional to µ.
When tested in water at 20°C flowing at 2 m/s, an 8-cmdiameter sphere has a measured drag of 5 N. What will be
the velocity and drag force on a 1.5-m-diameter weather
balloon moored in sea-level standard air under dynamically
similar conditions?
An automobile has a characteristic length and area of 8 ft
and 60 ft2, respectively. When tested in sea-level standard
air, it has the following measured drag force versus speed:
V, mi/h
20 40 60
Drag, lbf
31
115
249
The same car travels in Colorado at 65 mi/h at an altitude of
3500 m. Using dimensional analysis, estimate (a) its drag
force and (b) the horsepower required to overcome air drag.
The disk-gap-band parachute in the chapter-opener photo
had a drag of 1600 lbf when tested at 15 mi/h in air at 20°C
and 1 atm. (a) What was its drag coefficient? (b) If, as
stated, the drag on Mars is 65,000 lbf and the velocity is
375 mi/h in the thin Mars atmosphere, ρ ≈ 0.020 kg/m3,
what is the drag coefficient on Mars? (c) Can you explain
the difference between (a) and (b)?
Example 1.8 illustrates the fact that surface tension can
cause a fluid interface to rise or fall in a capillary tube. A
Problems 335
P5.8
P5.9
fluid property of importance in this example is surface
tension (Y), which has dimensions of force per unit
length. Show the dimensions of surface tension in terms
of basic dimensions.
The Archimedes number, Ar, used in the flow of stratified fluids, is a dimensionless combination of gravity g,
density difference Δρ, fluid width L, and viscosity μ.
Find the form of this number if it is proportional to g.
The Richardson number, Ri, which correlates the production of turbulence by buoyancy, is a dimensionless combination of the acceleration of gravity g, the fluid
temperature T0, the local temperature gradient ∂T/∂z,
and the local ­velocity gradient ∂u/∂z. Determine the
form of the Richardson number if it is proportional to g.
The principle of dimensional homogeneity
P5.10 Determine the dimension {MLTΘ} of the following
quantities:
2
∂u
∂ 2T
(b)
(p − p0 ) dA (c) ρcp
∂x
∂x ∂y
1
∂u
(d) e e e ρ
dx dy dz
∂t
(a) ρu
∫
All quantities have their standard meanings; for example,
ρ is density.
P5.11 During World War II, Sir Geoffrey Taylor, a British fluid
dynamicist, used dimensional analysis to estimate the
wave speed of an atomic bomb explosion. He assumed
that the blast wave radius R was a function of energy released E, air density ρ, and time t. Use dimensional reasoning to show how wave radius must vary with time.
P5.12 The Stokes number, St, used in particle dynamics studies, is a dimensionless combination of five variables:
acceleration of gravity g, viscosity µ, density ρ, particle
velocity U, and particle diameter D. (a) If St is proportional to µ and inversely proportional to g, find its form.
(b) Show that St is actually the quotient of two more
traditional ­dimensionless groups.
P5.13 The speed of propagation C of a capillary wave in deep
water is known to be a function only of density ρ,
wavelength λ, and surface tension Υ. Find the proper
functional relationship, completing it with a dimensionless constant. For a given density and wavelength,
how does the propagation speed change if the surface
tension is doubled?
The pi theorem or Ipsen’s method
P5.14 Flow in a pipe is often measured with an orifice plate, as
in Fig. P5.14. The volume flow Q is a function of the
pressure drop Δp across the plate, the fluid density ρ, the
pipe diameter D, and the orifice diameter d. Rewrite this
functional relationship in dimensionless form.
P5.14
P5.15 The wall shear stress τw in a boundary layer is assumed
to be a function of stream velocity U, boundary layer
thickness δ, local turbulence velocity u′, density ρ, and
local pressure gradient dp/dx. Using (ρ, U, δ) as repeating variables, rewrite this relationship as a dimensionless function.
P5.16 Convection heat transfer data are often reported as a heat
transfer coefficient h, defined by

Q = hA ΔT
.
where Q = heat flow, J/s
A = surface area, m2
ΔT = temperature difference, K
The dimensionless form of h, called the Stanton number,
is a combination of h, fluid density ρ, specific heat cp,
and flow velocity V. Derive the Stanton number if it is
proportional to h. What are the units of h?
P5.17 If you disturb a tank of length L and water depth h, the
surface will oscillate back and forth at frequency Ω,
­assumed here to depend also upon water density ρ and the
acceleration of gravity g. (a) Rewrite this as a dimensionless function. (b) If a tank of water sloshes at 2.0 Hz on
earth, how fast would it oscillate on Mars (g ≈ 3.7 m/s2)?
P5.18 Under laminar conditions, the volume flow Q through a
small triangular-section pore of side length b and length
L is a function of viscosity µ, pressure drop per unit
length Δp/L, and b. Using the pi theorem, rewrite this
relation in dimensionless form. How does the volume
flow change if the pore size b is doubled?
P5.19 The period of oscillation T of a water surface wave is
­assumed to be a function of density ρ, wavelength l, depth
h, gravity g, and surface tension Υ. Rewrite this relationship in dimensionless form. What results if Υ is negligible? Hint: Take l, ρ, and g as repeating variables.
336
Chapter 5 Dimensional Analysis and Similarity
P5.20 A fixed cylinder of diameter D and length L, immersed
in a stream flowing normal to its axis at velocity U, will
experience zero average lift. However, if the cylinder is
rotating at angular velocity Ω, a lift force F will arise.
The fluid density ρ is important, but viscosity is secondary and can be neglected. Formulate this lift behavior as
a dimensionless function.
P5.21 In Example 5.1 we used the pi theorem to develop Eq.
(5.2) from Eq. (5.1). Instead of merely listing the primary
­dimensions of each variable, some workers list the powers
of each primary dimension for each variable in an array:
F
M 1
L C 1
T −2
P5.22
P5.23
P5.24
P5.25
P5.26
L U ρ
0 0 1
1 1 −3
0 −1 0
μ
1
−1S
−1
This array of exponents is called the dimensional matrix
for the given function. Show that the rank of this matrix
(the size of the largest nonzero determinant) is equal to j
= n − k, the desired reduction between original variables
and the pi groups. This is a general property of dimensional matrices, as noted by Buckingham [1].
As will be discussed in Chap. 11, the power P developed by
a wind turbine is a function of diameter D, air density ρ,
wind speed V, and rotation rate ω. Viscosity effects are negligible. Rewrite this relationship in dimensionless form.
The period T of vibration of a beam is a function of its
length L, area moment of inertia I, modulus of elasticity
E, density ρ, and Poisson’s ratio σ. Rewrite this relation
in dimensionless form. What further reduction can we
make if E and I can occur only in the product form EI?
Hint: Take L, ρ, and E as repeating variables.
The lift force F on a missile is a function of its length L,
velocity V, diameter D, angle of attack α, density ρ, viscosity µ, and speed of sound a of the air. Write out the
dimensional matrix of this function and determine its
rank. (See Prob. P5.21 for an explanation of this concept.) Rewrite the function in terms of pi groups.
The thrust F of a propeller is generally thought to be a
function of its diameter D and angular velocity Ω, the forward speed V, and the density ρ and viscosity µ of the fluid.
Rewrite this relationship as a dimensionless function.
A pendulum has an oscillation period T which is assumed to depend on its length L, bob mass m, angle of
swing θ, and the acceleration of gravity. A pendulum 1 m
long, with a bob mass of 200 g, is tested on earth and
found to have a period of 2.04 s when swinging at 20°.
(a) What is its period when it swings at 45°? A similarly
constructed pendulum, with L = 30 cm and m = 100 g, is
to swing on the moon (g = 1.62 m/s2) at θ = 20°. (b)
What will be its period?
P5.27 In studying sand transport by ocean waves, A. Shields in
1936 postulated that the threshold wave-induced bottom
shear stress τ required to move particles depends on
gravity g, particle size d and density ρp, and water density ρ and viscosity µ. Find suitable dimensionless
groups of this problem, which resulted in 1936 in the
celebrated Shields sand transport diagram.
P5.28 A simply supported beam of diameter D, length L, and
modulus of elasticity E is subjected to a fluid crossflow of
velocity V, density ρ, and viscosity µ. Its center deflection
δ is assumed to be a function of all these variables. (a)
Rewrite this proposed function in dimensionless form. (b)
Suppose it is known that δ is independent of µ, inversely
proportional to E, and dependent only on ρV2, not ρ and V
separately. Simplify the dimensionless function accordingly. Hint: Take L, ρ, and V as repeating variables.
P5.29 When fluid in a pipe is accelerated linearly from rest,
it begins as laminar flow and then undergoes transition to turbulence at a time ttr that depends on the
pipe diameter D, fluid acceleration a, density ρ, and
viscosity µ. Arrange this into a dimensionless relation between ttr and D.
P5.30 When a large tank of high-pressure gas discharges

through a nozzle, the exit mass flow m is a function of
tank pressure p0 and temperature T0, gas constant R, specific heat cp, and nozzle diameter D. Rewrite this as a
dimensionless function. Check to see if you can use (p0,
T0, R, D) as ­repeating variables.
P5.31 The pressure drop per unit length in horizontal pipe
flow, Δp/L, depends on the fluid density ρ, viscosity µ,
diameter D, and volume flow rate Q. Rewrite this function in terms of pi groups.
P5.32 A weir is an obstruction in a channel flow that can be
calibrated to measure the flow rate, as in Fig. P5.32. The
volume flow Q varies with gravity g, weir width b into the
paper, and upstream water height H above the weir crest.
If it is known that Q is proportional to b, use the pi theorem to find a unique functional relationship Q(g, b, H).
H
Q
Weir
P5.32
P5.33 A spar buoy (see Prob. P2.113) has a period T of vertical
(heave) oscillation that depends on the waterline cross-
Problems 337
sectional area A, buoy mass m, and fluid specific weight
γ. How does the period change due to doubling of (a) the
mass and (b) the area? Instrument buoys should have
long periods to avoid wave resonance. Sketch a possible
long-period buoy design.
P5.34 To good approximation, the thermal conductivity k of a
gas (see Ref. 21 of Chap. 1) depends only on the density
ρ, mean free path l, gas constant R, and absolute temperature T. For air at 20°C and 1 atm, k ≈ 0.026 W/(m ·
K) and l ≈ 6.5 E–8 m. Use this information to determine
k for hydrogen at 20°C and 1 atm if l ≈ 1.2 E–7 m.
P5.35 The torque M required to turn the cone-plate viscometer
in Fig. P5.35 depends on the radius R, rotation rate Ω,
fluid viscosity µ, and cone angle θ. Rewrite this relation
in ­dimensionless form. How does the relation simplify it
if it is known that M is proportional to θ?
Ω
Nondimensionalizing the basic equations
R
θ
θ
Fluid
P5.35
b. If b = 120 ft and H = 15 in., the flow rate is 600 ft3/s.
What will be the flow rate if H = 3 ft?
P5.40 The time td to drain a liquid from a hole in the bottom of
a tank is a function of the hole diameter d, the initial
fluid volume υ0, the initial liquid depth h0, and the density ρ and viscosity µ of the fluid. Rewrite this relation as
a dimensionless function, using Ipsen’s method.
P5.41 A certain axial flow turbine has an output torque M that is
proportional to the volume flow rate Q and also depends
on the density ρ, rotor diameter D, and rotation rate Ω.
How does the torque change due to a doubling of (a) D
and (b) Ω?
P5.42 When disturbed, a floating buoy will bob up and down at
frequency f. Assume that this frequency varies with buoy
mass m, waterline diameter d, and the specific weight γ
of the liquid. (a) Express this as a dimensionless function. (b) If d and γ are constant and the buoy mass is
halved, how will the frequency change?
.
P5.36 The rate of heat loss Q loss through a window or wall is
a function of the temperature difference between inside and outside ΔT, the window surface area A, and
the R value of the window, which has units of (ft2 · h ·
°F)/ Btu. (a) Using the Buckingham Pi Theorem, find
an ­expression for rate of heat loss as a function of the
other three ­parameters in the problem. (b) If the temperature ­difference ΔT doubles, by what factor does
the rate of heat loss increase?
P5.37 The volume flow Q through an orifice plate is a function
of pipe diameter D, pressure drop Δp across the orifice,
fluid density ρ and viscosity µ, and orifice diameter d.
Using D, ρ, and Δp as repeating variables, express this
relationship in dimensionless form.
P5.38 The size d of droplets produced by a liquid spray nozzle
is thought to depend on the nozzle diameter D, jet velocity U, and the properties of the liquid ρ, µ, and Υ. Rewrite this relation in dimensionless form. Hint: Take D,
ρ, and U as repeating variables.
P5.39 The volume flow Q over a certain dam is a function of
dam width b, gravity g, and the upstream water depth H
above the dam crest. It is known that Q is proportional to
P5.43 Nondimensionalize the energy equation (4.75) and its
boundary conditions (4.62), (4.63), and (4.70) by defining T* = T/T0, where T0 is the inlet temperature, assumed
constant. Use other dimensionless variables as needed
from Eq. (5.11). Isolate all dimensionless parameters
you find, and relate them to the list given in Table 5.2.
P5.44 The differential energy equation for incompressible twodimensional flow through a “Darcy-type” porous medium is approximately
ρcp
σ ∂p ∂T
σ ∂p ∂T
∂ 2T
+ ρcp
+k 2 =0
μ ∂x ∂x
μ ∂y ∂y
∂y
where σ is the permeability of the porous medium. All
other symbols have their usual meanings. (a) What are
the appropriate dimensions for σ? (b) Nondimensionalize this equation, using (L, U, ρ, T0) as scaling constants,
and discuss any dimensionless parameters that arise.
P5.45 A model differential equation, for chemical reaction
­dynamics in a plug reactor, is as follows:
u
∂C
∂ 2C
∂C
= D 2 − kC −
∂x
∂t
∂x
where u is the velocity, D is a diffusion coefficient, k is a
reaction rate, x is distance along the reactor, and C is the
(dimensionless) concentration of a given chemical in the
reactor. (a) Determine the appropriate dimensions of D
and k. (b) Using a characteristic length scale L and average
­velocity V as parameters, rewrite this equation in dimensionless form and comment on any pi groups appearing.
P5.46 If a vertical wall at temperature Tw is surrounded by a
fluid at temperature T0, a natural convection boundary
338
Chapter 5 Dimensional Analysis and Similarity
layer flow will form. For laminar flow, the momentum
equation is
ρ(u
∂u
∂u
∂ 2u
+ υ ) = ρβ(T − T0 )g + μ 2
∂x
∂y
∂y
to be solved, along with continuity and energy, for (u, v,
T) with appropriate boundary conditions. The quantity β
is the thermal expansion coefficient of the fluid. Use ρ, g,
L, and (Tw−T0) to nondimensionalize this equation. Note
that there is no “stream” velocity in this type of flow.
P5.47 The differential equation for small-amplitude vibrations
y(x, t) of a simple beam is given by
P5.53
ρA
∂ 2y
∂t 2
+ EI
∂ 4y
∂x4
=0
where ρ = beam material density
A = cross-sectional area
I = area moment of inertia
E = Young’s modulus
P5.54
Use only the quantities ρ, E, and A to nondimensionalize y,
x, and t, and rewrite the differential equation in dimensionless form. Do any parameters remain? Could they be
­removed by further manipulation of the variables?
Data for spheres, cylinders, other bodies
P5.48 A smooth steel (SG = 7.86) sphere is immersed in a
stream of ethanol at 20°C moving at 1.5 m/s. Estimate its
drag in N from Fig. 5.2a. What stream velocity would
quadruple its drag? Take D = 2.5 cm.
P5.49 The sphere in Prob. P5.48 is dropped in gasoline at 20°C.
Ignoring its acceleration phase, what will its terminal
(constant) fall velocity be, from Fig. 5.2a?
P5.50 The parachute in the chapter-opener photo is, of course,
meant to decelerate the payload on Mars. The wind tunnel
test gave a drag coefficient of about 1.1, based upon the
projected area of the parachute. Suppose it was falling on
earth and, at an altitude of 1000 m, showed a steady descent
rate of about 18 mi/h. Estimate the weight of the payload.
P5.51 A ship is towing a sonar array that approximates a submerged cylinder 1 ft in diameter and 30 ft long with its
axis normal to the direction of tow. If the tow speed is 12
kn (1 kn = 1.69 ft/s), estimate the horsepower ­required
to tow this cylinder. What will be the frequency of vortices shed from the cylinder? Use Figs. 5.1 and 5.2.
P5.52 When fluid in a long pipe starts up from rest at a uniform
acceleration a, the initial flow is laminar. The flow undergoes transition to turbulence at a time t* which depends, to first approximation, only upon a, ρ, and µ.
Experiments by P. J. Lefebvre, on water at 20°C starting
from rest with 1-g acceleration in a 3-cm-diameter pipe,
P5.55
P5.56
showed transition at t* = 1.02 s. Use these data to estimate (a) the transition time and (b) the transition Reynolds number ReD for water flow accelerating at 35 m/s2
in a 5-cm-diameter pipe.
Vortex shedding can be used to design a vortex flowmeter (Fig. 6.34). A blunt rod stretched across the pipe
sheds vortices whose frequency is read by the sensor
downstream. Suppose the pipe diameter is 5 cm and the
rod is a cylinder of diameter 8 mm. If the sensor reads
5400 counts per minute, estimate the volume flow rate of
water in m3/h. How might the meter react to other liquids?
A fishnet is made of 1-mm-diameter strings knotted into
2 × 2 cm squares. Estimate the horsepower required to
tow 300 ft2 of this netting at 3 kn in seawater at 20°C.
The net plane is normal to the flow direction.
The radio antenna on a car begins to vibrate wildly at 8
Hz when the car is driven at 45 mi/h over a rutted road
that approximates a sine wave of amplitude 2 cm and
wavelength λ = 2.5 m. The antenna diameter is 4 mm. Is
the vibration due to the road or to vortex shedding?
Flow past a long cylinder of square cross-section results
in more drag than the comparable round cylinder. Here
are data taken in a water tunnel for a square cylinder of
side length b = 2 cm:
V, m/s
1.0
2.0
3.0
4.0
Drag, N/(m of depth)
21
85
191
335
(a) Use these data to predict the drag force per unit depth
of wind blowing at 6 m/s, in air at 20°C, over a tall
square chimney of side length b = 55 cm. (b) Is there any
uncertainty in your estimate?
P5.57 The simply supported 1040 carbon-steel rod of Fig.
P5.57 is subjected to a crossflow stream of air at 20°C
and 1 atm. For what stream velocity U will the rod center
deflection be approximately 1 cm?
D = 1 cm, L = 60 cm
δ = 1 cm?
U
P5.57 P5.58 For the steel rod of Prob. P5.57, at what airstream velocity
U will the rod begin to vibrate laterally in resonance in
its first mode (a half sine wave)? Hint: Consult a vibration text [31,32] under “lateral beam vibration.”
Problems 339
P5.59 A long, slender, smooth 3-cm-diameter flagpole bends
alarmingly in 20 mi/h sea-level winds, causing patriotic
citizens to gasp. An engineer claims that the pole will *P5.63
bend less if its surface is deliberately roughened. Is she
correct, at least qualitatively?
Scaling of model data
*P5.60 The thrust F of a free propeller, either aircraft or marine,
­depends upon density ρ, the rotation rate n in r/s, the diameter D, and the forward velocity V. Viscous effects are
slight and neglected here. Tests of a 25-cm-diameter
model aircraft propeller, in a sea-level wind tunnel, yield
the following thrust data at a velocity of 20 m/s:
Rotation rate, r/min
4800
6000
8000
Measured thrust, N
6.1
19
47
r
we
Po
Pump data
(Ω in r/s)
0
3Q
P
≈ 0.5 +
ΩD 3
ρ Ω3D 5
( (
▵p
Q
≈ 6.0 – 120
ρ Ω2D 2
ΩD 3
ω
M
(a) Use these data to make a crude but effective dimensionless plot. (b) Use the dimensionless data to predict
the thrust, in newtons, of a similar 1.6-m-diameter prototype propeller when rotating at 3800 r/min and flying at
225 mi/h at 4000-m standard altitude.
P5.61 If viscosity is neglected, typical pump flow results from
Example 5.3 are shown in Fig. P5.61 for a model pump
tested in water. The pressure rise decreases and the power
required increases with the dimensionless flow coefficient.
Curve-fit expressions are given for the data. Suppose a
similar pump of 12-cm diameter is built to move gasoline at
20°C and a flow rate of 25 m3/h. If the pump rotation speed
is 30 r/s, find (a) the pressure rise and (b) the power required.
Pressur
e ris
e
the rotation rate, and (b) the power delivered by the prototype. Assume sea-level air density.
The Keystone Pipeline in the Chap. 6 opener photo has
D = 36 in. and an oil flow rate Q = 590,000 barrels per day
(1 barrel = 42 U.S. gallons). Its pressure drop per unit
length, Δp/L, depends on the fluid density ρ, viscosity µ,
diameter D, and flow rate Q. A water-flow model test, at
20°C, uses a 5-cm-diameter pipe and yields Δp/L ≈ 4000
Pa/m. For dynamic similarity, estimate Δp/L of the pipeline. For the oil take ρ = 860 kg/m3 and µ = 0.005 kg/m . s.
P5.64 The natural frequency ω of vibration of a mass M attached to a rod, as in Fig. P5.64, depends only on M
2
Q
= flow coefficient
ΩD 3
P5.61
P5.62 For the system of Prob. P5.22, assume that a small model
wind turbine of diameter 90 cm, rotating at 1200 r/min,
delivers 280 watts when subjected to a wind of 12 m/s.
The data are to be used for a prototype of diameter 50 m
and winds of 8 m/s. For dynamic similarity, estimate (a)
L
Stiffness EI
P5.64 and the stiffness EI and length L of the rod. Tests with a 2-kg
mass attached to a 1040 carbon-steel rod of diameter 12 mm
and length 40 cm reveal a natural frequency of 0.9 Hz. Use
these data to predict the natural frequency of a 1-kg mass
attached to a 2024 aluminum alloy rod of the same size.
P5.65 In turbulent flow near a flat wall, the local velocity u
varies only with distance y from the wall, wall shear
stress τw, and fluid properties ρ and µ. The following
data were taken in the University of Rhode Island wind
tunnel for airflow, ρ = 0.0023 slug/ft3, µ = 3.81 E–7
slug/(ft · s), and τw = 0.029 lbf/ft2:
y, in
0.021
0.035
0.055
0.080
0.12
0.16
u, ft/s
50.6
54.2
57.6
59.7
63.5
65.9
(a) Plot these data in the form of dimensionless u versus
dimensionless y, and suggest a suitable power-law curve
fit. (b) Suppose that the tunnel speed is increased until u
= 90 ft/s at y = 0.11 in. Estimate the new wall shear
stress, in lbf/ft2.
P5.66 A torpedo 8 m below the surface in 20°C seawater cavitates at a speed of 21 m/s when atmospheric pressure is
101 kPa. If Reynolds number and Froude number effects
are negligible, at what speed will it cavitate when
­running at a depth of 20 m? At what depth should it be
to avoid cavitation at 30 m/s?
340
Chapter 5 Dimensional Analysis and Similarity
P5.67 A student needs to measure the drag on a prototype of
characteristic dimension dp moving at velocity Up in air
at standard atmospheric conditions. He constructs a
model of characteristic dimension dm, such that the ratio
dp/dm is some factor f. He then measures the drag on the
model at dynamically similar conditions (also with air at
standard atmospheric conditions). The student claims
that the drag force on the prototype will be identical to
that measured on the model. Is this claim correct? Explain.
P5.68 For the rotating-cylinder function of Prob. P5.20, if L >>
D, the problem can be reduced to only two groups, F/
(ρU2LD) versus (ΩD/U). Here are experimental data for
a cylinder 30 cm in diameter and 2 m long, rotating in
sea-level air, with U = 25 m/s.
Ω, rev/min
0
3000
6000
9000
12000 15000
F, N
0
850
2260
2900
3120
3300
(a) Reduce these data to the two dimensionless groups
and make a plot. (b) Use this plot to predict the lift of a
cylinder with D = 5 cm, L = 80 cm, rotating at 3800 rev/
min in water at U = 4 m/s.
P5.69 A simple flow measurement device for streams and
channels is a notch, of angle α, cut into the side of a dam,
as shown in Fig. P5.69. The volume flow Q depends
only on α, the acceleration of gravity g, and the height δ
of the ­upstream water surface above the notch vertex.
Tests of a model notch, of angle α = 55°, yield the following flow rate data:
δ, cm
10
20
30
40
3
8
47
126
263
Q, m /h
(a) Find a dimensionless correlation for the data. (b) Use
the model data to predict the flow rate of a prototype
notch, also of angle α = 55°, when the upstream height δ
is 3.2 m.
α
δ
V, ft/s
30
38
48
56
61
F, 1bf
1.25
1.95
3.02
4.05
4.81
Use these data to predict the drag force of a similar 15-in
diamond placed at similar orientation in 20°C water
flowing at 2.2 m/s.
P5.71 The pressure drop in a venturi meter (Fig. P3.128) varies
only with the fluid density, pipe approach velocity, and diameter ratio of the meter. A model venturi meter tested in
water at 20°C shows a 5-kPa drop when the approach velocity is 4 m/s. A geometrically similar prototype meter is
used to measure gasoline at 20°C and a flow rate of 9 m3/
min. If the prototype pressure gage is most accurate at 15
kPa, what should the upstream pipe diameter be?
P5.72 A one-twelfth-scale model of a large commercial aircraft
is tested in a wind tunnel at 20°C and 1 atm. The model
chord length is 27 cm, and its wing area is 0.63 m2. Test
results for the drag of the model are as follows:
V, mi/h
50
75
100
125
Drag, N
15
32
53
80
In the spirit of Fig. 5.7, use these data to estimate the
drag of the full-scale aircraft when flying at 550 mi/h, for
the same angle of attack, at 32,800 ft standard altitude.
P5.73 The power P generated by a certain windmill design depends on its diameter D, the air density ρ, the wind velocity V, the rotation rate Ω, and the number of blades n.
(a) Write this relationship in dimensionless form. A
model windmill, of diameter 50 cm, develops 2.7 kW at
sea level when V = 40 m/s and when rotating at 4800 r/
min. (b) What power will be developed by a geometrically and dynamically similar prototype, of diameter 5
m, in winds of 12 m/s at 2000 m standard altitude? (c)
What is the appropriate rotation rate of the prototype?
P5.74 A one-tenth-scale model of a supersonic wing tested at
700 m/s in air at 20°C and 1 atm shows a pitching
­moment of 0.25 kN · m. If Reynolds number effects are
negligible, what will the pitching moment of the prototype wing be if it is flying at the same Mach number at
8-km standard altitude?
Froude and Mach number scaling
P5.75 According to the website USGS Daily Water Data for
the Nation, the mean flow rate in the New River near
Hinton, WV, is 10,100 ft3/s. If the hydraulic model in
Fig. 5.8 is to match this condition with Froude number
P5.69
scaling, what is the proper model flow rate?
*P5.76 A 2-ft-long model of a ship is tested in a freshwater tow
P5.70 A diamond-shaped body, of characteristic length 9 in,
tank. The measured drag may be split into “friction”
has the following measured drag forces when placed in a
drag (Reynolds scaling) and “wave” drag (Froude scalwind tunnel at sea-level standard conditions:
ing). The model data are as follows:
Problems 341
Tow speed, ft/s
0.8
1.6
2.4
3.2
4.0
4.8
Friction drag, lbf
0.016
0.057
0.122
0.208
0.315
0.441
Wave drag, lbf
0.002
0.021
0.083
0.253
0.509
0.697
P5.77
P5.78
P5.79
P5.80
P5.81
P5.82
P5.83
P5.84
3-m height. If a one-fifteenth-scale model is tested in a
wave channel, what current speed, wave period, and
wave height should be encountered by the model?
Inventive rescaling of the data
The prototype ship is 150 ft long. Estimate its total drag
*P5.85 As shown in Example 5.3, pump performance data can
when cruising at 15 kn in seawater at 20°C.
3
be nondimensionalized. Problem P5.61 gave typical
A dam 75 ft wide, with a nominal flow rate of 260 ft , is
­dimensionless data for centrifugal pump “head,” H =
to be studied with a scale model 3 ft wide, using Froude
Δp/ρg, as follows:
scaling. (a) What is the expected flow rate for the model?
(b) What is the danger of only using Froude scaling for
gH
Q 2
≈ 6.0 − 120 ( 3 )
this test? (c) Derive a formula for a force on the model as
2 2
nD
nD
compared to a force on the prototype.
A prototype spillway has a characteristic velocity of 3 where Q is the volume flow rate, n the rotation rate in r/s,
m/s and a characteristic length of 10 m. A small model is
and D the impeller diameter. This type of correlation
constructed by using Froude scaling. What is the mini­allows one to compute H when (ρ, Q, D) are known. (a)
mum scale ratio of the model that will ensure that its
Show how to rearrange these pi groups so that one can
minimum Weber number is 100? Both flows use water at
size the pump, that is, compute D directly when (Q, H, n)
20°C.
are known. (b) Make a crude but effective plot of your
An East Coast estuary has a tidal period of 12.42 h (the
new function. (c) Apply part (b) to the following examsemidiurnal lunar tide) and tidal currents of approxiple: Find D when H = 37 m, Q = 0.14 m3/s, and n = 35
mately 80 cm/s. If a one-five-hundredth-scale model is
r/s. Find the pump diameter for this condition.
constructed with tides driven by a pump and storage ap- P5.86 Solve Prob. P5.49 for glycerin at 20°C, using the modiparatus, what should the period of the model tides be and
fied sphere-drag plot of Fig. 5.10.
what model current speeds are expected?
P5.87 In Prob. P5.61 it would be difficult to solve for Ω beA prototype ship is 35 m long and designed to cruise at
cause it appears in all three of the dimensionless pump
11 m/s (about 21 kn). Its drag is to be simulated by a
coefficients. Suppose that, in Prob. 5.61, Ω is unknown
1-m-long model pulled in a tow tank. For Froude scaling
but D = 12 cm and Q = 25 m3/h. The fluid is gasoline at
find (a) the tow speed, (b) the ratio of prototype to model
20°C. Rescale the coefficients, using the data of Prob.
drag, and (c) the ratio of prototype to model power.
P5.61, to make a plot of dimensionless power versus diAn airplane, of overall length 55 ft, is designed to fly at
mensionless rotation speed. Enter this plot to find the
680 m/s at 8000-m standard altitude. A one-thirtiethmaximum rotation speed Ω for which the power will not
scale model is to be tested in a pressurized helium wind
exceed 300 W.
tunnel at 20°C. What is the appropriate tunnel pressure P5.88 Modify Prob. P5.61 as follows: Let Ω = 32 r/s and Q =
in atm? Even at this (high) pressure, exact dynamic sim24 m3/h for a geometrically similar pump. What is the
ilarity is not achieved. Why?
maximum diameter if the power is not to exceed 340 W?
A one-fiftieth-scale model of a military airplane is tested
Solve this problem by rescaling the data of Fig. P5.61 to
at 1020 m/s in a wind tunnel at sea-level conditions. The
make a plot of dimensionless power versus dimensionmodel wing area is 180 cm2. The angle of attack is 3°. If
less diameter. Enter this plot directly to find the desired
the measured model lift is 860 N, what is the prototype
diameter.
lift, using Mach number scaling, when it flies at 10,000 P5.89 Wall friction τw, for turbulent flow at velocity U in a pipe
m standard altitude under dynamically similar condiof diameter D, was correlated, in 1911, with a
tions? Note: Be careful with the area scaling.
­dimensionless correlation by Ludwig Prandtl’s student
A one-fortieth-scale model of a ship’s propeller is tested
H. Blasius:
in a tow tank at 1200 r/min and exhibits a power output
τw
0.632
≈
of 1.4 ft · lbf/s. According to Froude scaling laws, what
2
ρU
(ρUD/μ) 1/4
should the revolutions per minute and horsepower output
of the prototype propeller be under dynamically similar Suppose that (ρ, D, µ, τw) were all known and it was desired to find the unknown velocity U. Rearrange and reconditions?
write the formula so that U can be immediately
A prototype ocean platform piling is expected to encouncalculated.
ter currents of 150 cm/s and waves of 12-s period and
342
Chapter 5 Dimensional Analysis and Similarity
P5.90 Knowing that Δp is proportional to L, rescale the data of
Example 5.10 to plot dimensionless Δp versus dimensionless viscosity. Use this plot to find the viscosity required in the first row of data in Example 5.10 if the
pressure drop is increased to 10 kPa for the same flow
rate, length, and density.
*P5.91 The traditional “Moody-type” pipe friction correlation
in Chap. 6 is of the form
f=
2ΔpD
2
ρV L
= fcn (
ρVD ε
,
μ D)
where D is the pipe diameter, L the pipe length, and ε the
wall roughness. Note that pipe average velocity V is
used on both sides. This form is meant to find Δp when
V is known. (a) Suppose that Δp is known, and we
wish to find V. Rearrange the above function so that V
is isolated on the left-hand side. Use the following
data, for ε/D = 0.005, to make a plot of your new function, with your velocity p­ arameter as the ordinate of
the plot.
f
pVD/µ
0.0356
0.0316
0.0308
0.0305
0.0304
15,000
75,000
250,000
900,000
3,330,000
(b) Use your plot to determine V, in m/s, for the following pipe flow: D = 5 cm, ε = 0.025 cm, L = 10 m, for
water flow at 20°C and 1 atm. The pressure drop Δp is
110 kPa.
Word Problems
W5.1 In 98 percent of data analysis cases, the “reducing factor” j, which lowers the number n of dimensional variables to n − j dimensionless groups, exactly equals the
number of relevant dimensions (M, L, T, Θ). In one case
(Example 5.5) this was not so. Explain in words why this
situation happens.
W5.2 Consider the following equation: 1 dollar bill ≈ 6 in. Is
this relation dimensionally inconsistent? Does it satisfy
the PDH? Why?
W5.3 In making a dimensional analysis, what rules do you follow for choosing your scaling variables?
W5.4 This chapter discusses the difficulty of scaling Mach and
Reynolds numbers together (an airplane) and Froude and
Reynolds numbers together (a ship). Give an example of
a flow that would combine Mach and Froude numbers.
Would there be scaling problems for common fluids?
W5.5 What is different about a very small model of a weir or
dam (Fig. P5.32) that would make the test results difficult to relate to the prototype?
W5.6 What else are you studying this term? Give an example
of a popular equation or formula from another course
(thermodynamics, strength of materials, or the like) that
does not satisfy the principle of dimensional homogeneity. ­Explain what is wrong and whether it can be modified to be homogeneous.
W5.7 Some colleges (such as Colorado State University) have
environmental wind tunnels that can be used to study
phenomena like wind flow over city buildings. What details of scaling might be important in such studies?
W5.8 If the model scale ratio is α = Lm/Lp, as in Eq. (5.19),
and the Weber number is important, how must the
model and p­ rototype surface tension be related to α for
dynamic similarity?
W5.9 For a typical incompressible velocity potential analysis
in Chap. 8 we solve ∇2ϕ = 0, subject to known values of
∂ϕ/∂n on the boundaries. What dimensionless parameters govern this type of motion?
Fundamentals of Engineering Exam Problems
FE5.1 Given the parameters (U, L, g, ρ, µ) that affect a certain
liquid flow problem, the ratio V2/(Lg) is usually known
as the
(a) velocity head, (b) Bernoulli head, (c) Froude number,
(d) kinetic energy, (e) impact energy
FE5.2 A ship 150 m long, designed to cruise at 18 kn, is to be
tested in a tow tank with a model 3 m long. The appropriate tow velocity is
(a) 0.19 m/s, (b) 0.35 m/s, (c) 1.31 m/s, (d) 2.55 m/s,
(e) 8.35 m/s
FE5.3 A ship 150 m long, designed to cruise at 18 kn, is to be
tested in a tow tank with a model 3 m long. If the model
wave drag is 2.2 N, the estimated full-size ship wave
drag is
(a) 5500 N, (b) 8700 N, (c) 38,900 N,
(d) 61,800 N, (e) 275,000 N
FE5.4 A tidal estuary is dominated by the semidiurnal lunar tide,
with a period of 12.42 h. If a 1:500 model of the estuary is
tested, what should be the model tidal period?
(a) 4.0 s, (b) 1.5 min, (c) 17 min, (d) 33 min, (e) 64 min
Comprehensive Problems 343
FE5.5 A football, meant to be thrown at 60 mi/h in sea-level air
(ρ = 1.22 kg/m3, µ = 1.78 E–5 N · s/m2), is to be tested
using a one-quarter scale model in a water tunnel (ρ =
998 kg/m3, µ = 0.0010 N · s/m2). For dynamic similarity,
what is the proper model water velocity?
(a) 7.5 mi/h, (b) 15.0 mi/h, (c) 15.6 mi/h,
(d) 16.5 mi/h, (e) 30 mi/h
FE5.6 A football, meant to be thrown at 60 mi/h in sea-level
air (ρ = 1.22 kg/m3, µ = 1.78 E–5 N · m2), is to be
tested ­using a one-quarter scale model in a water tunnel (ρ = 998 kg/m3, µ = 0.0010 N · s/m2). For dynamic
similarity, what is the ratio of prototype force to model
force?
(a) 3.86:1, (b) 16:1, (c) 32:1, (d) 56:1, (e) 64:1
FE5.7 Consider liquid flow of density ρ, viscosity µ, and velocity U over a very small model spillway of length scale L,
such that the liquid surface tension coefficient Y is important. The quantity ρU2L/Y in this case is important
and is called the
(a) capillary rise, (b) Froude number, (c) Prandtl number, (d) Weber number, (e) Bond number
FE5.8 If a stream flowing at velocity U past a body of length
L causes a force F on the body that depends only on U, L,
and fluid viscosity µ, then F must be proportional to
(a) ρUL/µ, (b) ρU2L2, (c) µU/L, (d) µUL, (e) UL/µ
FE5.9 In supersonic wind tunnel testing, if different gases are
used, dynamic similarity requires that the model and prototype have the same Mach number and the same
(a) Euler number, (b) speed of sound, (c) stagnation enthalpy, (d ) Froude number, (e) specific-heat ratio
FE5.10 The Reynolds number for a 1-ft-diameter sphere moving
at 2.3 mi/h through seawater (specific gravity 1.027, viscosity 1.07 E–3 N · s/m2) is approximately
(a) 300, (b) 3000, (c) 30,000, (d) 300,000,
(e) 3,000,000
FE5.11 The Ekman number, important in physical oceanography, is a dimensionless combination of µ, L, ρ, and the
earth’s rotation rate Ω. If the Ekman number is proportional to Ω, it should take the form
(a) ρΩ2L2/μ, (b) μΩL/ρ, (c) ρΩL/μ, (d) ρΩL2/μ,
(e) ρΩ/Lμ
FE5.12 A valid, but probably useless, dimensionless group is given
by (μT0g)/(ΥLα), where everything has its usual meaning, except α. What are the dimensions of α?
(a) ΘL−1T−1, (b) ΘL−1T−2, (c) ΘML−1, (d) Θ −1LT−1,
(e) ΘLT−1
Comprehensive Problems
C5.1
C5.2
Estimating pipe wall friction is one of the most common
tasks in fluids engineering. For long circular rough pipes in
turbulent flow, wall shear τw is a function of density
ρ, ­viscosity µ, average velocity V, pipe diameter d, and wall
roughness height e. Thus, functionally, we can write τw =
fcn(ρ, µ, V, d, e). (a) Using dimensional analysis, rewrite this
function in dimensionless form. (b) A certain pipe has d = 5
cm and ε = 0.25 mm. For flow of water at 20°C, measurements show the following values of wall shear stress:
Q, gal/min
1.5
3.0
6.0
9.0
12.0
14.0
τw, Pa
0.05
0.18
0.37
0.64
0.86
1.25
Plot these data using the dimensionless form obtained in
part (a) and suggest a curve-fit formula. Does your plot
reveal the entire functional relation obtained in part (a)?
When the fluid exiting a nozzle, as in Fig. P3.49, is a gas,
instead of water, compressibility may be important, especially if upstream pressure p1 is large and exit diameter
d2 is small. In this case, the difference p1 − p2 is no lon
ger controlling, and the gas mass flow m reaches a maximum value that depends on p1 and d2 and also on the
absolute upstream temperature T1 and the gas constant R.
Thus, functionally, ṁ = fcn(p1, d2, T1, R). (a) Using dimensional analysis, rewrite this function in dimensionless form. (b) A certain pipe has d2 = 1 cm. For flow of
air, measurements show the following values of mass
flow through the nozzle:
C5.3
C5.4
T1, K
300
300
300
500
800
p1, kPa
200
250
300
300
300
ṁ, kg/s
0.037
0.046
0.055
0.043
0.034
Plot these data in the dimensionless form obtained in
part (a). Does your plot reveal the entire functional relation obtained in part (a)?
Reconsider the fully developed draining vertical oil film
problem (see Fig. P4.80) as an exercise in dimensional
analysis. Let the vertical velocity be a function only of
­distance from the plate, fluid properties, gravity, and film
thickness. That is, w = fcn(x, ρ, µ, g, δ ). (a) Use
the pi ­
theorem to rewrite this function in terms of
­dimensionless parameters. (b) Verify that the exact solution
from Prob. P4.80 is consistent with your result in part (a).
The Taco Inc. model 4013 centrifugal pump has an impeller of diameter D = 12.95 in. When pumping 20°C water
at Ω = 1160 r/min, the measured flow rate Q and pressure
rise Δp are given by the manufacturer as follows:
Q, gal/min
600
700
Δp, lb/in2 36 35 34 32 29
200
300
400
500
23
344
Chapter 5 Dimensional Analysis and Similarity
(a) Assuming that Δp = fcn(ρ, Q, D, Ω), use the pi theorem to rewrite this function in terms of dimensionless
parameters and then plot the given data in dimensionless
form. (b) It is desired to use the same pump, running at
900 r/min, to pump 20°C gasoline at 400 gal/min. According to your dimensionless correlation, what pressure
rise Δp is ­expected, in lbf/in2?
C5.5 Does an automobile radio antenna vibrate in resonance
due to vortex shedding? Consider an antenna of length L
and diameter D. According to beam vibration theory [see
[31] or [32, p. 401]], the first mode natural frequency of
a solid circular cantilever beam is ωn = 3.516[EI/
(ρAL4)]1/2, where E is the modulus of elasticity, I is the
area moment of inertia, ρ is the beam material density,
and A is the beam cross-section area. (a) Show that ωn is
proportional to the antenna radius R. (b) If the antenna is
steel, with L = 60 cm and D = 4 mm, estimate the natural
vibration frequency, in Hz. (c) Compare with the shedding frequency if the car moves at 65 mi/h.
Design Projects
l = 0.212
l = 0.322
U, ft/s Ω, r/min U, ft/s
18.95
22.20
25.90
29.94
38.45
D5.2
roughened spheres [30] are given in Fig. D5.2. The
f­igure also shows typical golf ball data. We see that some
roughened spheres are better than golf balls in some regions. For the present study, let us neglect the ball’s spin,
which causes the very important side-force or Magnus
effect (see Fig. 8.15) and assume that the ball is hit without spin and follows the equations of motion for plane
motion (x, z):


mx = −F cos θ
mz = −F sin θ − W

ρπ
z


where F = CD D2 (x2 + z2 )
θ = tan−1 
24
x
We are given laboratory data, taken by Prof. Robert
Kirchhoff and his students at the University of Massachusetts, for the spin rate of a 2-cup anemometer. The anemometer was made of ping-pong balls (d = 1.5 in) split in
half, facing in opposite directions, and glued to thin (14 -in)
rods pegged to a center axle. (See Fig. P7.91 for a sketch.)
There were four rods, of lengths l = 0.212, 0.322, 0.458,
and 0.574 ft. The experimental data, for wind tunnel velocity U and rotation rate Ω, are as follows:
435
545
650
760
970
18.95
23.19
29.15
32.79
38.45
l = 0.458
l = 0.574
Ω, r/min U, ft/s Ω, r/min U, ft/s Ω, r/min
225
290
370
425
495
20.10
26.77
31.37
36.05
39.03
140
215
260
295
327
23.21
27.60
32.07
36.05
39.60
115
145
175
195
215
Assume that the angular velocity Ω of the device is a
function of wind speed U, air density ρ and viscosity µ,
rod length l, and cup diameter d. For all data, assume air
is at 1 atm and 20°C. Define appropriate pi groups for this
­problem, and plot the data in this dimensionless manner.
Comment on the possible uncertainty of the results.
As a design application, suppose we are to use this
­anemometer geometry for a large-scale (d = 30 cm) airport wind anemometer. If wind speeds vary up to 25 m/s
and we desire an average rotation rate Ω = 120 r/min,
what should be the proper rod length? What are possible
limitations of your design? Predict the expected Ω (in r/
min) of your d­ esign as affected by wind speeds from 0 to
25 m/s.
By analogy with the cylinder drag data in Fig. 5.2b,
spheres also show a strong roughness effect on drag, at
least in the Reynolds number range 4 E4 < ReD < 3 E5,
which ­accounts for the dimpling of golf balls to increase
their distance traveled. Some experimental data for
The ball has a particular CD(ReD) curve from Fig. D5.2 and
is struck with an initial velocity V0 and angle θ0. Take the
ball’s average mass to be 46 g and its diameter to be 4.3
cm. ­Assuming sea-level air and a modest but finite range
of initial conditions, integrate the equations of motion to
compare the trajectory of “roughened spheres” to actual
golf ball calculations. Can the rough sphere outdrive a
normal golf ball for any conditions? What roughness-effect differences occur between a low-impact duffer and,
say, Tiger Woods?
0.6
Drag coefficient, CD
D5.1
Golf ball
0.5
0.4
0.3
0.2
900 ×10–5
1250 ×10–5
500 ×10–5
0.1
∈ = 150 × 10–5
D
0
2 ×104
D5.2
Rough spheres
Smooth sphere
105
106
Reynolds number, UD/ν
4 ×106
References 345
References
1.
2.
3.
4.
5.
6.
7.
8.
9.
10.
11.
12.
13.
14.
15.
16.
17.
18.
19.
E. Buckingham, “On Physically Similar Systems: Illustrations of the Use of Dimensional Equations,” Phys. Rev.,
vol. 4, no. 4, 1914, pp. 345–376.
J. D. Anderson, Computational Fluid Dynamics: The Basics with Applications, McGraw-Hill, New York, 1995.
P. W. Bridgman, Dimensional Analysis, Yale University
Press, New Haven, CT, 1922, rev. ed., 1963.
H. L. Langhaar, Dimensional Analysis and the Theory of
Models, Wiley, New York, 1951.
E. C. Ipsen, Units, Dimensions, and Dimensionless Numbers, McGraw-Hill, New York, 1960.
H. G. Hornung, Dimensional Analysis: Examples of the
Use of Symmetry, Dover, New York, 2006.
E. S. Taylor, Dimensional Analysis for Engineers, Clarendon Press, Oxford, England, 1974.
G. I. Barenblatt, Dimensional Analysis, Gordon and
Breach, New York, 1987.
A. C. Palmer, Dimensional Analysis and Intelligent Experimentation, World Scientific Publishing, Hackensack, NJ,
2008.
T. Szirtes, Applied Dimensional Analysis and Modeling, 2d
ed., Butterworth-Heinemann, Burlington, MA, 2006.
R. Esnault-Pelterie, Dimensional Analysis and Metrology,
F. Rouge, Lausanne, Switzerland, 1950.
R. Kurth, Dimensional Analysis and Group Theory in Astrophysics, Pergamon, New York, 1972.
R. Kimball and M. Ross, The Data Warehouse Toolkit: The
Complete Guide to Dimensional Modeling, 2d ed., Wiley,
New York, 2002.
R. Nakon, Chemical Problem Solving Using Dimensional
Analysis, Prentice-Hall, Upper Saddle River, NJ, 1990.
D. R. Maidment (ed.), Hydrologic and Hydraulic Modeling
Support: With Geographic Information Systems, Environmental Systems Research Institute, Redlands, CA, 2000.
A. M. Curren, Dimensional Analysis for Meds, 4th ed.,
­Delmar Cengage Learning, Independence, KY, 2009.
G. P. Craig, Clinical Calculations Made Easy: Solving
­Problems Using Dimensional Analysis, 4th ed., Lippincott
Williams and Wilkins, Baltimore, MD, 2008.
M. Zlokarnik, Dimensional Analysis and Scale-Up in
­Chemical Engineering, Springer-Verlag, New York, 1991.
W. G. Jacoby, Data Theory and Dimensional Analysis,
Sage, Newbury Park, CA, 1991.
20.
21.
22.
23.
24.
25.
26.
27.
28.
29.
30.
31.
32.
33.
34.
35.
36.
37.
B. Schepartz, Dimensional Analysis in the Biomedical
­Sciences, Thomas, Springfield, IL, 1980.
T. Horntvedt, Calculating Dosages Safely: A Dimensional
Analysis Approach, F. A. Davis Co., Philadelphia, PA,
2012.
J. B. Bassingthwaighte et al., Fractal Physiology, Oxford
Univ. Press, New York, 1994.
K. J. Niklas, Plant Allometry: The Scaling of Form and
­Process, Univ. of Chicago Press, Chicago, 1994.
A. Roshko, “On the Development of Turbulent Wakes from
Vortex Streets,” NACA Rep. 1191, 1954.
G. W. Jones, Jr., “Unsteady Lift Forces Generated by Vortex Shedding about a Large, Stationary, Oscillating Cylinder at High Reynolds Numbers,” ASME Symp. Unsteady
Flow, 1968.
J. Kunes, Dimensionless Physical Quantities in Science
and Engineering, Elsevier, New York, 2012.
V. P. Singh et al. (eds.), Hydraulic Modeling, Water
­Resources Publications LLC, Highlands Ranch, CO, 1999.
L. Armstrong, Hydraulic Modeling and GIS, ESRI Press,
La Vergne, TN, 2011.
R. Ettema, Hydraulic Modeling: Concepts and Practice,
American Society of Civil Engineers, Reston, VA, 2000.
R. D. Blevins, Applied Fluid Dynamics Handbook, van
­Nostrand Reinhold, New York, 1984.
W. J. Palm III, Mechanical Vibration, Wiley, New York,
2006.
S. S. Rao, Mechanical Vibrations, 5th ed., Prentice-Hall,
­Upper Saddle River, NJ, 2010.
J. B. Barlow, W. H. Rae, and A. Pope, Low-Speed Wind
­Tunnel Testing, Wiley, New York, 1999.
B. H. Goethert, Transonic Wind Tunnel Testing, Dover,
New York, 2007.
American Institute of Aeronautics and Astronautics, Recommended Practice: Wind Tunnel Testing, 2 vols., Reston,
VA, 2003.
P. N. Desai, J. T. Schofield, and M. E. Lisano, “Flight Reconstruction of the Mars Pathfinder Disk-Gap-Band Parachute Drag Coefficients,” J. Spacecraft Rockets, vol. 42,
no. 4, July–August 2005, pp. 672–676.
K.-H. Kim, “Recent Advances in Cavitation Research,”
14th International Symposium on Transport Phenomena,
­Honolulu, HI, March 2012.
This chapter is mostly about flow analysis. The photo shows the 36-inch-diameter K
­ eystone
Pipeline, which has been operating since July 2010. This pipeline delivers heavy and light
blends of oil from Hardisty, Alberta, Canada, to refineries in Texas. The pipeline is
­completely buried except for where it emerges occasionally at delivery or pump stations;
it can currently deliver up to 700,000 barrels of oil per day. [Image courtesy of ­TransCanada]
346
Chapter 6
Viscous Flow in Ducts
Motivation. From Chap. 3 to Chap. 5 we have prepared the basic analytic tools
for fluid dynamics. The applications of fluid flow are classified as internal or
external, depending on whether the fluid is forced to flow in a duct or over a
body. A majority of practical engineering problems for fluids involve internal and
external flows, but these two kinds of flows exhibit very different characteristics.
This chapter is completely devoted to fluid flow in ducts, with various velocities,
various fluids, and various duct shapes.
The basic piping problem is this: Given the pipe geometry and its added components (such as fittings, valves, bends, and diffusers) plus the desired flow rate
and fluid properties, what pressure drop is needed to drive the flow? Of course,
it may be stated in alternative form: Given the pressure drop available from a
pump, what flow rate will ensue? The correlations discussed in this chapter are
adequate to solve most such piping problems.
This chapter is for incompressible flow; Chap. 9 treats compressible pipe flow.
6.1 Reynolds Number Regimes
Now that we have derived and studied the basic flow equations in Chap. 4, you
would think that we could just whip off myriad beautiful solutions illustrating the
full range of fluid behavior, of course expressing all these educational results in
dimensionless form, using our new tool from Chap. 5, dimensional analysis.
The fact of the matter is that no general analysis of fluid motion yet exists. There
are several dozen known particular solutions, there are many approximate digital
computer solutions, and there are a great many experimental data. There is a lot of
theory available if we neglect such important effects as viscosity and compressibility (Chap. 8), but there is no general theory and there may never be. The reason
is that a profound and vexing change in fluid behavior occurs at moderate Reynolds
numbers. The flow ceases being smooth and steady (laminar) and becomes fluctuating and agitated (turbulent). The changeover is called transition to turbulence.
In Fig. 5.2a we saw that transition on the cylinder and sphere occurred at about
Re = 3 × 105, where the sharp drop in the drag coefficient appeared. Transition
347
348
Chapter 6 Viscous Flow in Ducts
u
Fig. 6.1 The three regimes of viscous flow: (a) laminar flow at low
Re; (b) transition at intermediate
Re; (c) turbulent flow at high Re.
u
Small natural
disturbances
damp quickly
(a)
u
Continuous
turbulence
Intermittent
bursts of
turbulence
t
(b)
t
(c)
t
depends on many effects, such as wall roughness (Fig. 5.2b) or fluctuations in the
inlet stream, but the primary parameter is the Reynolds number. There are a great
many data on transition but only a small amount of theory [1 to 3].
Turbulence can be detected from a measurement by a small, sensitive instrument such as a hot-wire anemometer (Fig. 6.29e) or a piezoelectric pressure
transducer. The flow will appear steady on average but will reveal rapid, random
fluctuations if turbulence is present, as sketched in Fig. 6.1. If the flow is laminar,
there may be occasional natural disturbances that damp out quickly (Fig. 6.1a).
If transition is occurring, there will be sharp bursts of intermittent turbulent fluctuation (Fig. 6.1b) as the increasing Reynolds number causes a breakdown or
instability of laminar motion. At sufficiently large Re, the flow will fluctuate
continually (Fig. 6.1c) and is termed fully turbulent. The fluctuations, typically
ranging from 1 to 20 percent of the average velocity, are not strictly periodic but
are random and encompass a continuous range, or spectrum, of frequencies. In a
typical wind tunnel flow at high Re, the turbulent frequency ranges from 1 to
10,000 Hz, and the wavelength ranges from about 0.01 to 400 cm.
EXAMPLE 6.1
The accepted transition Reynolds number for flow in a circular pipe is Red,crit ≈ 2300.
For flow through a 5-cm-diameter pipe, at what velocity will this occur at 20°C for (a)
airflow and (b) water flow?
Solution
Almost all pipe flow formulas are based on the average velocity V = Q/A, not centerline
or any other point velocity. Thus transition is specified at ρVd/µ ≈ 2300. With d known,
we introduce the appropriate fluid properties at 20°C from Tables A.3 and A.4:
(a) Air:
ρVd (1.205 kg/m3 )V(0.05 m)
= 2300 or
=
μ
1.80 E−5 kg/(m · s)
(b) Water:
ρVd (998 kg/m3 )V(0.05 m)
= 2300 or
=
μ
0.001 kg/(m · s)
V ≈ 0.7
m
s
V = 0.046
m
s
These are very low velocities, so most engineering air and water pipe flows are turbulent, not laminar. We might expect laminar duct flow with more viscous fluids such
as lubricating oils or glycerin.
6.1 Reynolds Number Regimes 349
Fig. 6.2 Flow issuing at constant
speed from a faucet: (a) highviscosity, low-Reynolds-number,
laminar flow; (b) low-viscosity,
high-Reynolds-number, turbulent
flow. Note the ragged, disorderly
shape of the jet.
(a) Laminar flow
(b) Turbulent flow
In free-surface flows, turbulence can be observed directly. Figure 6.2 shows
water flow issuing from a faucet. This phenomenon can be easily observed in our
daily life. The low-Reynolds-number jet (Fig. 6.2a) is smooth and laminar, with the
fast center motion and slower wall flow forming different ­trajectories joined by a
liquid sheet. The higher-Reynolds-number turbulent flow (Fig. 6.2b) is unsteady
and irregular but, when averaged over time, is steady and predictable.
How did turbulence form inside the pipe? The laminar parabolic flow profile, which
is similar to Eq. (4.141), became unstable as early as Red reaches about 1000, began
to form “slugs” or “puffs.” A puff has a fast-moving front and a slow-moving rear and
may be visualized by experimenting with glass tube flow. Figure 6.3 shows water
flowing in a pipe of 1 cm in diameter at Reynolds number from 2000 to 15000, photographed by Dr. Tom Mullin and Dr. Jorge Peixinho. For Red = 2000 (Fig. 6.3a),
there is an irregular laminar-turbulent interface, and vortex roll-up visible near the
entrance. Although a puff swirls around, it does not constitute turbulent yet. It can be
considered as a seed of turbulence. Structurally, a puff has limited in space, and does
not spread out to the entire pipe. More importantly, it is limited in time. A puff will
be floating down the pipe, showing no signs of ill health, and disappear suddenly at
early transition period. For Red = 3000 and 4000, more puffs can be observed in the
pipe (Fig. 6.3b and c). When Reynolds number increases further to 7500 and 10,000,
Flow
(a)
Fig. 6.3 Formation of a turbulent
puff in pipe flow: (a) Red = 2000;
(b) Red = 3000; (c) Red = 4000;
(d) Red = 7500; (e) Red = 10,000.
(Courtesy of Dr. Thomas Mullin/
Dr. Jorge Peixinho)
Red = 2000
(b)
Red = 3000
(c)
Red = 4000
(d)
Red = 7500
(e)
Red = 10,000
350
Chapter 6 Viscous Flow in Ducts
it is believed that the birth rate of puffs far exceeds the death rate, turbulence will
spread to the entire pipe as shown in Fig. 6.3d and e.
A complete description of the statistical aspects of turbulence is given in
Ref. 1, while theory and data on transition effects are given in Refs. 2 and 3. At
this introductory level we merely point out that the primary parameter affecting
transition is the Reynolds number. If Re = UL/ν, where U is the average stream
velocity and L is the “width,” or transverse thickness, of the shear layer, the following approximate ranges occur:
0<
1<
100 <
103 <
104 <
106 <
Re
Re
Re
Re
Re
Re
< 1: highly viscous laminar “creeping” motion
< 100: laminar, strong Reynolds number dependence
< 103: laminar, boundary layer theory useful
< 104: transition to turbulence
< 106: turbulent, moderate Reynolds number dependence
< ∞: turbulent, slight Reynolds number dependence
These representative ranges vary somewhat with flow geometry, surface roughness, and the level of fluctuations in the inlet stream. The great majority of our
analyses are concerned with laminar flow or with turbulent flow, and one should
not normally design a flow operation in the transition region.
Historical Outline
Since turbulent flow is more prevalent than laminar flow, experimenters have
observed turbulence for centuries without being aware of the details. Before 1930
flow instruments were too insensitive to record rapid fluctuations, and workers
simply reported mean values of velocity, pressure, force, and so on. But turbulence can change the mean values dramatically, as with the sharp drop in drag
coefficient in Fig. 5.2. A German engineer named G. H. L. Hagen first reported
in 1839 that there might be two regimes of viscous flow. He measured water flow
in long brass pipes and deduced a pressure-drop law:
Δp = (const)
LQ
R4
+ entrance effect
(6.1)
This is exactly our laminar flow scaling law from Example 5.4, but Hagen did
not realize that the constant was proportional to the fluid viscosity.
The formula broke down as Hagen increased Q beyond a certain limit—that
is, past the critical Reynolds number—and he stated in his paper that there must
be a second mode of flow characterized by “strong movements of water for which
Δp varies as the second power of the discharge. . . .” He admitted that he could
not clarify the reasons for the change.
A typical example of Hagen’s data is shown in Fig. 6.4. The pressure drop
varies linearly with V = Q/A up to about 1.1 ft/s, where there is a sharp change.
Above about V = 2.2 ft/s the pressure drop is nearly quadratic with V. The actual
power Δp ∝ V1.75 seems impossible on dimensional grounds but is easily explained
when the dimensionless pipe flow data (Fig. 5.9) are displayed.
In 1883 Osborne Reynolds, a British engineering professor, showed that the
change depended on the parameter ρVd/µ, now named in his honor. By i­ ntroducing
6.1 Reynolds Number Regimes 351
120
Turbulent flow
▵p α V 1.75
100
Pressure drop ▵p, lbf/ft2
80
60
40
Laminar flow
▵p α V
Dye filament
20
Transition
Needle
Tank
(a)
0
0
0.5
1.0
1.5
Average velocity V, ft/s
2.0
2.5
Fig. 6.4 Experimental evidence of transition for water flow in a 14 -in
smooth pipe 10 ft long.
(b)
(c)
Fig. 6.5 Reynolds’ sketches of pipe
flow transition: (a) low-speed,
­laminar flow; (b) high-speed,
turbulent flow; (c) spark photograph
of condition (b).
Source: Reynolds, “An Experimental
­Investigation of the Circumstances which
Determine Whether the Motion of Water
Shall Be Direct or Sinuous and of the
Law of Resistance in Parallel C
­ hannels,”
Phil. Trans. R. Soc., vol. 174, 1883,
pp. 935–982.
a dye streak into a pipe flow, Reynolds could observe transition and turbulence.
His sketches [4] of the flow behavior are shown in Fig. 6.5.
If we examine Hagen’s data and compute the Reynolds number at V = 1.1 ft/s,
we obtain Red = 2100. The flow became fully turbulent, V = 2.2 ft/s, at Red =
4200. The accepted design value for pipe flow transition is now taken to be
Red,crit ≈ 2300
(6.2)
This is accurate for commercial pipes (Fig. 6.13), although with special care in
providing a rounded entrance, smooth walls, and a steady inlet stream, Red,crit can
be delayed until much higher values. The study of transition in pipe flow, both
experimentally and theoretically, continues to be a fascinating topic for researchers, as discussed in a recent review article [55]. Note: The value of 2300 is for
transition in pipes. Other geometries, such as plates, airfoils, cylinders, and
spheres, have completely different transition Reynolds numbers.
Transition also occurs in external flows around bodies such as the sphere and
cylinder in Fig. 5.2. Ludwig Prandtl, a German engineering professor, showed
in 1914 that the thin boundary layer surrounding the body was undergoing tran-
352
Chapter 6 Viscous Flow in Ducts
sition from laminar to turbulent flow. Thereafter the force coefficient of a body
was acknowledged to be a function of the Reynolds number [Eq. (5.2)].
There are now extensive theories and experiments of laminar flow instability
that explain why a flow changes to turbulence. Reference 5 is an advanced
­textbook on this subject.
Laminar flow theory is now well developed, and many solutions are known
[2, 3], but no analyses can simulate the fine-scale random fluctuations of turbulent flow.1 Therefore most turbulent flow theory is semiempirical, based on
dimensional analysis and physical reasoning; it is concerned with the mean flow
properties only and the mean of the fluctuations, not their rapid variations. The
turbulent flow “theory” presented here in Chaps. 6 and 7 is unbelievably crude
yet surprisingly effective. We shall attempt a rational approach that places
­turbulent flow analysis on a firm physical basis.
6.2 Internal Viscous Flows
An internal flow is constrained by the bounding walls, and the viscous effects
will grow and meet and permeate the entire flow. Figure 6.6 shows an internal
flow in a long duct. There is an entrance region where a nearly inviscid upstream
Growing
boundary
layers
Inviscid
core flow
Boundary
layers
merge
Developed
velocity
profile u(r)
r
x
u(r, x)
Entrance length Le
(developing profile region)
Fully developed
flow region
Pressure
Entrance
pressure
drop
Fig. 6.6 Developing velocity
­profiles and pressure changes in
the entrance of a duct flow.
Linear
pressure
drop in
fully developed
flow region
0
Le
x
1
However, direct numerical simulation (DNS) of low-Reynolds-number turbulence is now quite
­common [32].
6.2 Internal Viscous Flows 353
flow converges and enters the tube. Viscous boundary layers grow downstream,
retarding the axial flow u(r, x) at the wall and thereby accelerating the center
core flow to maintain the incompressible continuity requirement
∫
Q = u dA = const
(6.3)
At a finite distance from the entrance, the boundary layers merge and the
inviscid core disappears. The tube flow is then entirely viscous, and the axial
velocity adjusts slightly further until at x = Le it no longer changes with x and is
said to be fully developed, u ≈ u(r) only. Downstream of x = Le the velocity
profile is constant, the wall shear is constant, and the pressure drops linearly with
x, for either laminar or turbulent flow. All these details are shown in Fig. 6.6.
Dimensional analysis shows that the Reynolds number is the only parameter
­affecting entrance length. If
Q
Le = f (d, V, ρ, μ)
V=
A
ρVd
Le
then
= g(
= g(Red )
(6.4)
μ )
d
For laminar flow [2, 3], the accepted correlation is
Le
≈ 0.06 Red
laminar
d
(6.5)
The maximum laminar entrance length, at Red,crit = 2300, is Le = 138d, which is
the longest development length possible.
In turbulent flow, the boundary layers grow faster, and Le is relatively shorter.
For decades, the writer has favored a sixth-power-law estimate, Le/d ≈ 4.4 Re1/6
d ,
but recent CFD results, communicated by Fabien Anselmet, and separately by
Sukanta Dash, indicate that a better turbulent entrance–length correlation is
Le
≈ 1.6 Re1/4
d
d
for Red ≤ 107
(6.6)
Some computed turbulent entrance–length estimates are thus
Red
4000
104
105
106
107
Le/d
13
16
28
51
90
Now 90 diameters may seem “long,” but typical pipe flow applications involve
an L/d value of 1000 or more, in which case the entrance effect may be neglected
and a simple analysis made for fully developed flow. This is p­ ossible for both
laminar and turbulent flows, including rough walls and noncircular cross sections.
EXAMPLE 6.2
A 12 -in-diameter water pipe is 60 ft long and delivers water at 5 gal/min at 20°C. What
fraction of this pipe is taken up by the entrance region?
354
Chapter 6 Viscous Flow in Ducts
Solution
Convert
Q = (5 gal/min)
0.00223 ft3/s
= 0.0111 ft3/s
1 gal/min
The average velocity is
V=
Q
0.0111 ft3/s
=
= 8.17 ft/s
A (π/4) [ ( 12 /12) ft] 2
From Table 1.4 read for water ν = 1.01 × 10−6 m2/s = 1.09 × 10−5 ft2/s. Then the pipe
Reynolds number is
Red =
1
Vd (8.17 ft/s) [ ( 2 /12) ft]
=
= 31,300
ν
1.09 × 10−5 ft2/s
This is greater than 4000; hence the flow is fully turbulent, and Eq. (6.6) applies for
entrance length:
Le
1/4
≈ 1.6 Re1/4
= 21
d = (1.6) (31,300)
d
The actual pipe has L/d = (60 ft)/[(12 /12)ft] = 1440. Hence the entrance region takes
up the fraction
Le
21
=
= 0.015 = 1.5%
Ans.
L
1440
This is a very small percentage, so we can reasonably treat this pipe flow as essentially
fully developed.
Shortness can be a virtue in duct flow if one wishes to maintain the inviscid
core. For example, a “long” wind tunnel would be ridiculous, since the viscous core
would invalidate the purpose of simulating free-flight conditions. A typical laboratory low-speed wind tunnel test section is 1 m in diameter and 5 m long, with V =
30 m/s. If we take νair = 1.51 × 10−5 m2/s from Table 1.4, then Red = 1.99 × 106
and, from Eq. (6.6), Le/d ≈ 60. The test section has L/d = 5, which is much shorter
than the development length. At the end of the section the wall boundary layers are
only 10 cm thick, leaving 80 cm of inviscid core suitable for model testing.
As can be seen in the pressure changes in the entrance of a duct flow shown
in Fig. 6.6, the pressure drop in the entrance region is significantly higher than
that in the fully developed region. This is because the wall shear stress is the highest at the duct inlet where the thickness of the boundary layer is smallest. The wall
shear stress decreases gradually until the boundary layer becomes fully developed.
6.3 Head Loss—The Friction Factor
When applying pipe flow formulas to practical problems, it is customary to use
a control volume analysis. Consider incompressible steady flow between sections
6.3 Head Loss—The Friction Factor 355
1 p1 = p 2 + ∆p
g x = g sin ϕ
g
r=
ϕ
R
r
u(r
)
τw
τ(r
Z1
x2
–x
1
=
2 p2
)
ϕ
L
x
Fig. 6.7 Control volume, just inside
the pipe wall, of steady, fully
­developed flow between two
­sections in an inclined pipe.
Z2
1 and 2 of the inclined constant-area pipe in Fig. 6.7. The one-dimensional continuity ­relation, Eq. (3.30), reduces to
Q1 = Q2 = const
or
V1 = V2 = V
since the pipe is of constant area. The steady flow energy equation (3.75) becomes
p
p
V2
V2
+α
+α
+
z
=
+ z) + hf
( ρg
)1 ( ρg
2g
2g
2
(6.7)
since there is no pump or turbine between 1 and 2. For fully developed flow, the
velocity profile shape is the same at sections 1 and 2. Thus α1 = α2 and, since
V1 = V2, Eq. (6.7) reduces to head loss versus pressure drop and elevation change:
p2
Δp
p1
hf = (z1 − z2 ) + ( − ) = Δz +
ρg ρg
ρg
(6.8)
The pipe head loss equals the change in the sum of pressure and gravity head—
that is, the change in height of the hydraulic grade line (HGL).
Finally, apply the momentum relation (3.40) to the control volume in Fig. 6.7,
accounting for applied x-directed forces due to pressure, gravity, and shear:
∑ Fx = Δp (πR2 ) + ρg(πR2 )L sin ϕ − τw (2πR)L = m (V2 − V1 ) = 0 (6.9a)
Rearrange this and we find that the head loss is also related to wall shear stress:
Δz +
Δp
2τw L 4τw L
= hf =
=
ρg
ρg R
ρg d
(6.9b)
356
Chapter 6 Viscous Flow in Ducts
where we have substituted Δz = L sin ϕ from the geometry of Fig. 6.7. Note that,
regardless of whether the pipe is horizontal or tilted, the head loss is proportional
to the wall shear stress.
How should we correlate the head loss for pipe flow problems? The answer
was given a century and a half ago by Julius Weisbach, a German professor who
in 1850 published the first modern textbook on hydrodynamics. Equation (6.9b)
shows that hf is proportional to (L/d), and data such as Hagen’s in Fig. 6.6 show
that, for turbulent flow, hf is approximately proportional to V2. The proposed correlation, still as effective today as in 1850, is
hf = f
L V2
d 2g
ε
where f = fcn(Red, , duct shape) d
(6.10)
The dimensionless parameter f is called the Darcy friction factor, after Henry Darcy
(1803–1858), a French engineer whose pipe flow experiments in 1857 first ­established
the effect of roughness on pipe resistance. The quantity ε is the wall roughness height,
which is important in turbulent (but not laminar) pipe flow. We added the “duct
shape” effect in Eq. (6.10) to remind us that square and triangular and other noncircular ducts have a somewhat different friction factor than a circular pipe. Actual data
and theory for friction factors will be discussed in the sections that follow.
By equating Eqs. (6.9) and (6.10) we find an alternative form for friction factor:
f=
8τw
ρV 2
(6.11)
For noncircular ducts, we must interpret τw to be an average value around the
duct perimeter. For this reason Eq. (6.10) is preferred as a unified definition of
the Darcy friction factor.
6.4 Laminar Fully Developed Pipe Flow
Analytical solutions can be readily derived for laminar flows, either circular or
noncircular. Consider fully developed Poiseuille flow in a round pipe of diameter
d, radius R. Complete analytical results were given in Sec. 4.10. Let us review
those formulas here:
u = umax (1 −
dp R2
dp
Δp + ρgΔz
r2
where
u
=
−
and (− ) = (
max
2)
(
)
)
dx 4μ
dx
L
R
Δp + ρgΔz R2
Q umax
V= =
=(
) 8μ
A
2
L
∫
Q = udA = πR2V =
∣ ∣
τw = μ
du
dr
πR4 Δp + ρgΔz
)
8μ (
L
4μV 8μV R Δp + ρgΔz
=
= (
)
R
d
2
L
r=R
32μLV 128μLQ
hf =
=
ρgd2
πρgd 4
=
(6.12)
6.4 Laminar Fully Developed Pipe Flow 357
The paraboloid velocity profile has an average velocity V which is one-half of
the maximum velocity. The quantity Δp is the pressure drop in a pipe of length
L; that is, (dp/dx) is negative. These formulas are valid whenever the pipe
­Reynolds number, Red = ρVd/µ, is less than about 2300. Note that τw is proportional to V (see Fig. 6.6) and is independent of density because the fluid
­acceleration is zero. Neither of these is true in turbulent flow.
With wall shear stress known, the Poiseuille flow friction factor is easily determined:
flam =
8τw,lam
ρV
2
=
8(8μV/d)
ρV 2
=
64
64
=
ρVd/μ Red
(6.13)
In laminar flow, the pipe friction factor decreases inversely with Reynolds number. This famous formula is effective, but often the algebraic relations of Eqs.
(6.12) are more direct for problems.
EXAMPLE 6.3
An oil with ρ = 900 kg/m3 and ν = 0.0002 m2/s flows upward through an inclined
pipe as shown in Fig. E6.3. The pressure and elevation are known at sections 1 and
2, 10 m apart. Assuming steady laminar flow, (a) verify that the flow is up, (b) compute hf between 1 and 2, and compute (c) Q, (d) V, and (e) Red. Is the flow really
laminar?
Solution
Part (a)
For later use, calculate
μ = ρν = (900 kg/m3 ) (0.0002 m2/s) = 0.18 kg/(m · s)
z2 = ΔL sin 40° = (10 m) (0.643) = 6.43 m
d = 6 cm
2
10 m
1
E6.3
Q,V
40°
p1 = 350,000 Pa, z1 = 0
p2 = 250,000 Pa
358
Chapter 6 Viscous Flow in Ducts
The flow goes in the direction of falling HGL; therefore, compute the hydraulic gradeline height at each section:
HGL1 = z1 +
p1
350,000
=0+
= 39.65 m
ρg
900(9.807)
HGL2 = z2 +
p2
250,000
= 34.75 m
= 6.43 +
ρg
900(9.807)
The HGL is lower at section 2; hence the flow is up from 1 to 2 as assumed.
Ans. (a)
Part (b)
The head loss is the change in HGL:
hf = HGL1 − HGL2 = 39.65 m − 34.75 m = 4.9 m
Ans. (b)
Half the length of the pipe is quite a large head loss.
Part (c)
We can compute Q from the various laminar flow formulas, notably Eq. (6.12):
Q=
πρgd 4hf
128μL
=
π(900) (9.807) (0.06) 4 (4.9)
= 0.0076 m3/s
128(0.18) (10)
Ans. (c)
Part (d)
Divide Q by the pipe area to get the average velocity:
V=
0.0076
= 2.7 m/s
π(0.03) 2
Ans. (d )
Vd 2.7(0.06)
=
= 810
ν
0.0002
Ans. (e)
Q
πR
2
=
Part (e)
With V known, the Reynolds number is
Red =
This is well below the transition value Red = 2300, so we are fairly certain the flow is
laminar.
Notice that by sticking entirely to consistent SI units (meters, seconds, kilograms,
newtons) for all variables we avoid the need for any conversion factors in the
calculations.
EXAMPLE 6.4
A liquid of specific weight ρg = 58 lbf/ft3 flows by gravity through a 1-ft tank and a
1-ft capillary tube at a rate of 0.15 ft3/h, as shown in Fig. E6.4. Sections 1 and 2 are
at atmospheric pressure. Neglecting entrance effects and friction in the large tank,
compute the viscosity of the liquid.
6.5 Turbulence Modeling 359
1
Solution
∙ System sketch: Figure E6.4 shows L = 1 ft, d = 0.004 ft, and Q = 0.15 ft3/h.
∙ Assumptions: Laminar, fully developed, incompressible (Poiseuille) pipe flow.
Atmospheric pressure at sections 1 and 2. Negligible velocity at surface, V1 ≈ 0.
∙ Approach: Use continuity and energy to find the head loss and thence the viscosity.
∙ Property values: Given ρg = 58 lbf/ft3, figure out ρ = 58/32.2 = 1.80 slug/ft3 if
needed.
∙ Solution step 1: From continuity and the known flow rate, determine V2:
1 ft
1 ft
d = 0.004 ft
2
Q = 0.15
E6.4
V2 =
Q
Q
(0.15/3600)ft3/s
=
=
= 3.32 ft/s
2
A2 (π/4)d
(π/4) (0.004 ft) 2
Write the energy equation between 1 and 2, canceling terms, and find the head loss:
p1 α1V21
p2 α2V22
+
+ z1 =
+
+ z2 + hf
ρg
ρg
2g
2g
ft3/ h
or
hf = z1 − z2 −
α2V22
(2.0) (3.32 ft/s) 2
= 2.0 ft − 0 ft −
= 1.66 ft
2g
2(32.2 ft/s2 )
∙ Comment: We introduced α2 = 2.0 for laminar pipe flow from Eq. (3.76). If we
forgot α2, we would have calculated hf = 1.83 ft, a 10 percent error.
∙ Solution step 2: With head loss known, the viscosity follows from the laminar
formula in Eqs. (6.12):
hf = 1.66 ft =
32 μLV
(ρg)d2
=
32μ(1.0 ft) (3.32 ft/s)
(58 lbf/ft3 ) (0.004 ft) 2
solve for μ = 1.45 E-5
slug
ft-s
Ans.
∙ Comments: We didn’t need the value of ρ—the formula contains ρg, but who knew?
Note also that L in this formula is the pipe length of 1 ft, not the total elevation
change.
∙ Final check: Calculate the Reynolds number to see if it is less than 2300 for laminar
flow:
Red =
ρVd (1.80 slug/ft3 ) (3.32 ft/s) (0.004 ft)
≈ 1650 Yes, laminar.
=
μ
(1.45 E−5 slug/ft-s)
∙ Comments: So we did need ρ after all to calculate Red.
∙ Unexpected comment: For this head loss, there is a second (turbulent) solution, as
we shall see in Example 6.8.
6.5 Turbulence Modeling
Throughout this chapter we assume constant density and viscosity and no thermal
interaction, so that only the continuity and momentum equations are to be solved
for velocity and pressure
∂u ∂υ ∂w
+
+
=0
∂x ∂y
∂z
Continuity:
Momentum:
ρ
dV
= −∇p + ρg + μ ∇ 2V
dt
(6.14)
360
Chapter 6 Viscous Flow in Ducts
subject to no slip at the walls and known inlet and exit conditions. (We shall save
our free-surface solutions for Chap. 10.)
We will not work with the differential energy relation, Eq. (4.53), in this
chapter, but it is very important, both for heat transfer calculations and for general
understanding of duct flow processes. There is work being done by pressure
forces to drive the fluid through the duct. Where does this energy go? There is
no work done by the wall shear stresses, because the velocity at the wall is zero.
The answer is that pressure work is balanced by viscous dissipation in the interior
of the flow. The integral of the dissipation function Φ, from Eq. (4.50), over the
flow field will equal the pressure work. An example of this fundamental viscous
flow energy balance is given in Problem C6.7.
Both laminar and turbulent flows satisfy Eqs. (6.14). For laminar flow,
where there are no random fluctuations, we go right to the attack and solve
them for a variety of geometries [2, 3], leaving many more, of course, for the
problems.
Reynolds’ Time-Averaging Concept
For turbulent flow, because of the fluctuations, every velocity and pressure term
in Eqs. (6.14) is a rapidly varying random function of time and space. At present
our mathematics cannot handle such instantaneous fluctuating variables. No single pair of random functions V(x, y, z, t) and p(x, y, z, t) is known to be a solution to Eqs. (6.14). Moreover, our attention as engineers is toward the average or
mean values of velocity, pressure, shear stress, and the like in a high-Reynoldsnumber (turbulent) flow. This approach led Osborne Reynolds in 1895 to rewrite
Eqs. (6.14) in terms of mean or time-averaged turbulent variables.
The time mean u of a turbulent function u(x, y, z, t) is defined by
u=
1
T
T
∫ u dt
0
(6.15)
where T is an averaging period taken to be longer than any significant period of
the fluctuations themselves. The mean values of turbulent velocity and pressure
are illustrated in Fig. 6.8. For turbulent gas and water flows, an averaging period
T ≈ 5 s is usually quite adequate.
The fluctuation u′ is defined as the deviation of u from its average value
u′ = u − u
(6.16)
also shown in Fig. 6.8. It follows by definition that a fluctuation has zero mean
value:
u′ =
1
T
T
∫ (u − u) dt = u − u = 0
0
(6.17)
However, the mean square of a fluctuation is not zero and is a measure of the
intensity of the turbulence:
u′ 2 =
1
T
T
∫ u′
0
2
dt ≠ 0
(6.18)
6.5 Turbulence Modeling 361
u
p
p = p + p'
u = u + u'
u'
p
u
p'
Fig. 6.8 Definition of mean and
fluctuating turbulent variables:
(a) velocity; (b) pressure.
t
t
(a)
(b)
Nor in general are the mean fluctuation products such as u′υ′ and u′p′ zero in
a typical turbulent flow.
Reynolds’ idea was to split each property into mean plus fluctuating variables:
u = u + u′
υ = υ + υ′
w = w + w′
p = p + p′
(6.19)
Substitute these into Eqs. (6.14), and take the time mean of each equation. The
c­ ontinuity relation reduces to
∂u ∂υ ∂w
+
+
=0
∂x ∂y
∂z
(6.20)
which is no different from a laminar continuity relation.
However, each component of the momentum equation (6.14b), after time averaging, will contain mean values plus three mean products, or correlations, of
fluctuating velocities. The most important of these is the momentum relation in
the mainstream, or x, direction, which takes the form
ρ
∂p
du
∂
∂u
= − + ρgx +
μ − ρu′ 2)
dt
∂x
∂x ( ∂x
∂
∂u
∂
∂u
+
μ − ρu′υ ′ ) +
μ − ρu ′ w ′ ) ∂y ( ∂y
∂z ( ∂z
(6.21)
The three correlation terms −ρu′ 2, −ρu′υ′, and −ρu′w′ are called turbulent
stresses or Reynolds stresses because they have the same dimensions and occur
right alongside the newtonian (laminar) stress terms μ(∂u/∂x) and so on.
The turbulent stresses are unknown a priori and must be related by experiment
to geometry and flow conditions, as detailed in Refs. 1 to 3. Fortunately, in duct
and boundary layer flows, the stress −ρu′υ′, associated with direction y normal
362
Chapter 6 Viscous Flow in Ducts
to the wall, is dominant, and we can approximate with excellent accuracy a simpler streamwise momentum equation
ρ
where
∂p
du
∂τ
≈ − + ρgx +
dt
∂x
∂y
τ=μ
(6.22)
∂u
− ρu′υ′ = τlam + τturb
∂y
(6.23)
Figure 6.9 shows the distribution of τlam and τturb from typical measurements
across a turbulent shear layer near a wall. Laminar shear is dominant near the
wall (the narrow region adjacent to the wall is called viscous wall layer or sublayer), and turbulent shear dominates in the outer layer. There is an intermediate
region, called the overlap layer, where both laminar and turbulent shear are
important. These three regions are labeled in Fig. 6.9.
In the outer layer τturb is two or three orders of magnitude greater than τlam,
and vice versa in the viscous sublayer. These experimental facts enable us to use
a crude but very effective model for the velocity distribution u(y) across a turbulent wall layer.
The Logarithmic Overlap Law
We have seen in Fig. 6.9 that there are three regions in turbulent flow near a wall:
1. Viscous sublayer: Viscous shear dominates.
2. Outer layer: Turbulent shear dominates.
3. Overlap layer: Both types of shear are important.
From now on let us agree to drop the overbar from velocity u. Let τw be the wall
shear stress, and let δ and U represent the thickness and velocity at the edge of
the outer layer, y = δ.
For the viscous sublayer, Prandtl deduced in 1930 that u must be independent
of the shear layer thickness:
u = f (μ, τw, ρ, y)
y
(6.24)
y
y = δ (x)
U(x)
Outer
turbulent
layer
τ (x, y)
τ turb
Fig. 6.9 Typical velocity and shear
distributions in turbulent flow near a
wall: (a) shear; (b) velocity.
u(x, y)
Overlap layer
Viscous sublayer
τ lam
τ w(x)
(a)
0
(b)
6.5 Turbulence Modeling 363
By dimensional analysis, this is equivalent to
u+ =
yu*
u
=F(
ν )
u*
τw 1/2
u* = ( )
ρ
(6.25)
Equation (6.25) is called the law of the wall, and the quantity u* is termed the
friction velocity because it has dimensions {LT −1}, although it is not actually a
flow velocity.
Subsequently, Kármán in 1933 deduced that u in the outer layer is independent
of molecular viscosity, but its deviation from the stream velocity U must depend
on the layer thickness δ and the other properties:
(U − u) outer = g(δ, τw, ρ, y)
(6.26)
Again, by dimensional analysis we rewrite this as
y
U−u
=G( )
u*
δ
(6.27)
where u* has the same meaning as in Eq. (6.25). Equation (6.27) is called the
v­ elocity-defect law for the outer layer.
Both the wall law (6.25) and the defect law (6.27) are found to be accurate
for a wide variety of experimental turbulent duct and boundary layer flows [Refs.
1 to 3]. They are different in form, yet they must overlap smoothly in the intermediate layer. In 1937 C. B. Millikan showed that this can be true only if the
overlap layer velocity varies logarithmically with y:
u
1 yu*
+ B overlap layer
= ln
v
u* κ
(6.28)
Over the full range of turbulent smooth wall flows, the dimensionless constants
κ and B are found to have the approximate values κ ≈ 0.41 and B ≈ 5.0. Equation
(6.28) is called the logarithmic overlap layer.
Thus by dimensional reasoning and physical insight we infer that a plot of u
versus ln y in a turbulent shear layer will show a curved wall region, a curved
outer region, and a straight-line logarithmic overlap. Figure 6.10 shows that this
is exactly the case. The four outer-law profiles shown all merge smoothly with the
logarithmic overlap law but have different magnitudes because they vary in external pressure gradient. The wall law is unique and follows the linear viscous relation
u+ =
yu*
u
=
= y+
ν
u*
(6.29)
from the wall to about y+ = 5, thereafter curving over to merge with the logarithmic law at about y+ = 30.
Believe it or not, Fig. 6.10, which is nothing more than a shrewd correlation
of velocity profiles, is the basis for most existing “theory” of turbulent shear
flows. Notice that we have not solved any equations at all but have merely
expressed the streamwise velocity in a neat form.
364
Chapter 6 Viscous Flow in Ducts
30
Outer law profiles:
Strong increasing pressure
Flat plate flow
Pipe flow
Strong decreasing pressure
25
Linear
viscous
sublayer,
Eq. (6.29)
u+ = y +
u+ =
Ov
15
er
ay
l
lap
er
u
u*
20
Logarithmic
overlap
Eq. (6.28)
10
ne
r la
yer
Experimental data
In
5
Fig. 6.10 Experimental verification
of the inner, outer, and overlap layer
laws relating velocity profiles in
­turbulent wall flow.
0
1
10
10 2
yu*
v
10 3
10 4
y+ =
There is serendipity in Fig. 6.10: The logarithmic law (6.28), instead of just
being a short overlapping link, actually approximates nearly the entire velocity
profile, except for the outer law when the pressure is increasing strongly downstream (as in a diffuser). The inner wall law typically extends over less than 2
percent of the profile and can be neglected. Thus we can use Eq. (6.28) as an
excellent approximation to solve nearly every turbulent flow problem presented in
this and the next chapter. Many additional applications are given in Refs. 2 and 3.
Advanced Modeling Concepts
Turbulence modeling is a very active field. Scores of papers have been published
to more accurately simulate the turbulent stresses in Eq. (6.21) and their y and z
components. This research, now available in advanced texts [1, 13, 19], goes well
beyond the present book, which is confined to the use of the logarithmic law
(6.28) for pipe and boundary layer problems. For example, L. Prandtl, who
invented boundary layer theory in 1904, later proposed an eddy viscosity model
of the Reynolds stress term in Eq. (6.23):
−ρ u′v′ = τturb ≈ μt
du
dy
where
μt ≈ ρ l 2
∣ ∣
du
dy
(6.30)
The term µt, which is a property of the flow, not the fluid, is called the eddy
viscosity and can be modeled in various ways. The most popular form is Eq. (6.30),
6.5 Turbulence Modeling 365
where l is called the mixing length of the turbulent eddies (analogous to mean
free path in molecular theory). Near a solid wall, l is approximately proportional
to distance from the wall, and Kármán suggested
l ≈ κy where κ = Kármán’s constant ≈ 0.41
(6.31)
As a homework assignment, Prob. P6.40, you may show that Eqs. (6.30) and
(6.31) lead to the logarithmic law (6.28) near a wall.
Modern turbulence models approximate three-dimensional turbulent flows and
employ additional partial differential equations for such quantities as the turbulence kinetic energy, the turbulent dissipation, and the six Reynolds stresses. For
details, see Refs. 1, 13, and 19.
EXAMPLE 6.5
Air at 20°C flows through a 14-cm-diameter tube under fully developed conditions.
The centerline velocity is u0 = 5 m/s. Estimate from Fig. 6.10 (a) the friction velocity
u* and (b) the wall shear stress τw.
Solution
u ( y)
u0 = 5 m /s
y=R
r
r = R = 7 cm
y
∙ System sketch: Figure E6.5 shows turbulent pipe flow with u0 = 5 m/s and
R = 7 cm.
∙ Assumptions: Figure 6.10 shows that the logarithmic law, Eq. (6.28), is reasonable
all the way to the center of the tube.
∙ Approach: Use Eq. (6.28) to estimate the unknown friction velocity u*.
∙ Property values: For air at 20°C, ρ = 1.205 kg/m3 and ν = 1.51 E−5 m2/s.
∙ Solution step: Insert all the given data into Eq. (6.28) at y = R (the centerline). The
only unknown is u*:
u0 1
Ru*
= ln
+B
u* κ ( v )
E6.5
or
(0.07 m)u*
5.0 m/s
1
=
ln
+5
u*
0.41 [ 1.51 E−5 m2/s ]
Although the logarithm makes it awkward, one can solve this either by hand or by Excel
­iteration. There is an automatic iteration procedure in Excel—File, Excel Options, Formulas, Enable iterative calculation—but here we simply show how to iterate by repeat
calculations, copied and pasted downward. For a single unknown, in this case u*, we only
need two columns, one for the unknown and one for the equation. The writer hopes that
the following copy-and-iterate procedure is clear:
A
1
2
3
4
Here place the first guess for u*:
1.0
=B1 (the number, not the equation)
Copy A2 here
Keep copying down…
B
Here place the equation that solves for u*:
=(5.0/(1/0.41*ln(0.07*a1/1.51E−5)+5))
Copy B1 equation and place here
Copy B2 here
Keep copying down until convergence
366
Chapter 6 Viscous Flow in Ducts
Note that B2 uses the cell location for u*, A1, not the notation u*. Here are the actual
numbers, not instructions or equations, for this problem:
A
B
1.0000
0.1954
0.2314
0.2271
0.2275
0.1954
0.2314
0.2271
0.2275
0.2275
The solution for u* has converged to 0.2275. To three decimal places,
u* ≈ 0.228 m/s
Ans. (a)
τw = ρu*2 = (1.205) (0.228) 2 ≈ 0.062 Pa
Ans. (b)
∙ Comments: The logarithmic law solved everything! This is a powerful technique,
using an experimental velocity correlation to approximate general turbulent flows.
You may check that the Reynolds number Red is about 40,000, definitely turbulent
flow.
6.6 Turbulent Pipe Flow
For turbulent pipe flow we need not solve a differential equation but instead
proceed with the logarithmic law, as in Example 6.5. Assume that Eq. (6.28)
correlates the local mean velocity u(r) all the way across the pipe
u(r) 1 (R − r)u*
≈ ln
+B
κ
ν
u*
(6.32)
where we have replaced y with R − r. Compute the average velocity from this
profile:
V=
Q
1
= 2
A πR
=
∫ u*[ 1κ ln (R −νr)u* + B]2πr dr
R
0
1
2 Ru*
3
u*( ln
+ 2B − )
κ
ν
κ
2
(6.33)
Introducing κ = 0.41 and B = 5.0, we obtain, numerically,
V
Ru*
≈ 2.44 ln
+ 1.34
ν
u*
(6.34)
This looks only marginally interesting until we realize that V/u* is directly related
to the Darcy friction factor:
ρV 1/2
V
8 1/2
=(
=
(f)
τw )
u*
2
(6.35)
6.6 Turbulent Pipe Flow 367
Moreover, the argument of the logarithm in (6.34) is equivalent to
1
f 1/2
Ru* 2Vd u* 1
=
= Red ( )
ν
ν V
2
8
(6.36)
Introducing (6.35) and (6.36) into Eq. (6.34), changing to a base-10 logarithm,
and rearranging, we obtain
1
f 1/2
≈ 1.99 log (Red f 1/2 ) − 1.02
(6.37)
In other words, by simply computing the mean velocity from the logarithmic law
correlation, we obtain a relation between the friction factor and Reynolds number
for turbulent pipe flow. Prandtl derived Eq. (6.37) in 1935 and then adjusted the
constants slightly to fit friction data better:
1
f 1/2
= 2.0 log (Red f 1/2 ) − 0.8
(6.38)
This is the accepted formula for a smooth-walled pipe. Some numerical values
may be listed as follows:
Red
f
4000
104
105
106
107
108
0.0399
0.0309
0.0180
0.0116
0.0081
0.0059
Thus f drops by only a factor of 5 over a 10,000-fold increase in Reynolds number. Equation (6.38) is cumbersome to solve if Red is known and f is wanted.
There are many alternative approximations in the literature from which f can be
computed explicitly from Red:
0.316 Re−1/4
4000 < Red < 105
d
f= •
Red −2
1.8
log
(
6.9 )
H. Blasius (1911)
Ref. 9, Colebrook
(6.39)
However, Eq. (6.38), the preferred formula, is easily solved by computer iteration.
Blasius, a student of Prandtl, presented his formula in the first correlation ever
made of pipe friction versus Reynolds number. Although his formula has a limited
range, it illustrates what was happening in Fig. 6.4 to Hagen’s 1839 pressure-drop
data. For a horizontal pipe, from Eq. (6.39),
hf =
or
Δp
μ 1/4 L V2
L V2
=f
≈ 0.316 (
ρg
d 2g
ρVd ) d 2g
Δp ≈ 0.158 Lρ3/4μ1/4d−5/4V7/4
(6.40)
at low turbulent Reynolds numbers. This explains why Hagen’s data for pressure
drop begin to increase as the 1.75 power of the velocity, in Fig. 6.4. Note that
Δp varies only slightly with viscosity, which is characteristic of turbulent flow.
Introducing Q = 14πd2V into Eq. (6.40), we obtain the alternative form
Δp ≈ 0.241Lρ3/4μ1/4d−4.75Q1.75
(6.41)
368
Chapter 6 Viscous Flow in Ducts
For a given flow rate Q, the turbulent pressure drop decreases with diameter even
more sharply than the laminar formula (6.12). Thus the quickest way to reduce
required pumping pressure is to increase the pipe size, although, of course, the
larger pipe is more expensive. Doubling the pipe size decreases Δp by a factor
of about 27 for a given Q. Compare Eq. (6.40) with Example 5.7 and Fig. 5.9.
The maximum velocity in turbulent pipe flow is given by Eq. (6.32), evaluated
at r = 0:
umax 1 Ru*
≈ ln
+B
κ
ν
u*
(6.42)
Combining this with Eq. (6.33), we obtain the formula relating mean velocity to
maximum velocity:
V
≈ (1 + 1.3 √f ) −1
umax
(6.43)
Some numerical values are
Red
4000
104
105
106
107
108
V/umax
0.794
0.814
0.852
0.877
0.895
0.909
The ratio varies with the Reynolds number and is much larger than the value of
0.5 predicted for all laminar pipe flow in Eq. (6.12). Thus a turbulent velocity
profile, as shown in Fig. 6.11b, is very flat in the center and drops off sharply
to zero at the wall.
Effect of Wall Roughness
It was not known until experiments in 1800 by Coulomb [6] that surface
roughness has an effect on friction resistance. It turns out that the effect is
negligible for laminar pipe flow, and all the laminar formulas derived in this
section are valid for rough walls also. But turbulent flow is strongly affected
by roughness. In Fig. 6.10 the linear viscous sublayer extends out only to
umax
V
(a)
V
umax
Fig. 6.11 Comparison of laminar
and turbulent pipe flow velocity
profiles for the same volume flow:
(a) laminar flow; (b) turbulent flow.
(b)
Parabolic
curve
6.6 Turbulent Pipe Flow 369
0.10
0.08
∍
= 0.0333
d
0.06
0.0163
0.04
0.00833
f
0.00397
u
u*
0.02
oth
o
Sm
0.00198
64
Red
∆B
0.00099
h
ug
Ro
Eq. (6.39a)
Eq. (6.38)
≈1n∍ +
log
yu*
v
(a)
0.01
10 3
10 4
10 5
10 6
Red
(b)
Fig. 6.12 Effect of wall roughness on turbulent pipe flow. (a) The logarithmic overlap
­velocity profile shifts down and to the right; (b) experiments with sand-grain roughness by
Nikuradse [7] show a systematic increase of the turbulent friction factor with the roughness
ratio.
y+ = yu*/ν = 5. Thus, compared with the diameter, the sublayer thickness ys
is only
ys 5ν/u*
14.1
=
=
d
d
Red f 1/2
(6.44)
For example, at Red = 105, f = 0.0180, and ys/d = 0.001, a wall roughness of
about 0.001d will break up the sublayer and profoundly change the wall law in
Fig. 6.10.
Measurements of u(y) in turbulent rough-wall flow by Prandtl’s student
Nikuradse [7] show, as in Fig. 6.12a, that a roughness height ε will force the
logarithm law profile outward on the abscissa by an amount approximately
equal to ln ε+, where ε+ = εu*/ν. The slope of the logarithm law remains the
same, 1/κ, but the shift outward causes the constant B to be less by an amount
ΔB ≈ (1/κ) ln ε+.
Nikuradse [7] simulated roughness by gluing uniform sand grains onto the
inner walls of the pipes. He then measured the pressure drops and flow rates and
correlated friction factor versus Reynolds number in Fig. 6.12b. We see that
laminar friction is unaffected, but turbulent friction, after an onset point, increases
monotonically with the roughness ratio ε/d. For any given ε/d, the friction factor
370
Chapter 6 Viscous Flow in Ducts
becomes constant (fully rough) at high Reynolds numbers. These points of change
are certain values of ε+ = εu*/ν:
εu*
< 5: hydraulically smooth walls, no effect of roughness on
ν
friction
εu*
5≤
≤ 70: transitional roughness, moderate Reynolds number effect
ν
εu*
> 70: fully rough flow, sublayer totally broken up and friction
ν
independent of Reynolds number
For fully rough flow, ε+ > 70, the log law downshift ΔB in Fig. 6.12a is
1
ln ε+ − 3.5
κ
and the logarithm law modified for roughness becomes
ΔB ≈
(6.45)
1
1 y
ln y+ + B − ΔB = ln + 8.5
κ
κ ε
u+ =
(6.46)
The viscosity vanishes, and hence fully rough flow is independent of the Reynolds
number. If we integrate Eq. (6.46) to obtain the average velocity in the pipe, we obtain
V
d
= 2.44 ln + 3.2
ε
u*
1
ε/d
= −2.0 log
fully rough flow
1/2
3.7
f
or
(6.47)
There is no Reynolds number effect; hence the head loss varies exactly as the square
of the velocity in this case. Some numerical values of friction factor may be listed:
ε/d
0.00001
0.0001
0.001
0.01
0.05
f
0.00806
0.0120
0.0196
0.0379
0.0716
The friction factor increases by 9 times as the roughness increases by a factor of
5000. In the transitional roughness region, sand grains behave somewhat differently from commercially rough pipes, so Fig. 6.12b has now been replaced by
the Moody chart.
The Moody Chart
In 1939 to cover the transitionally rough range, Colebrook [9] combined the
smooth wall [Eq. (6.38)] and fully rough [Eq. (6.47)] relations into a clever
interpolation formula:
1
f
1/2
= −2.0 log (
ε/d
2.51
+
3.7 Red f 1/2 )
(6.48)
This is the accepted design formula for turbulent friction. It was plotted in 1944
by Moody [8] into what is now called the Moody chart for pipe friction (Fig. 6.13).
The Moody chart is probably the most famous and useful figure in fluid mechanics.
6.6 Turbulent Pipe Flow 371
Values of (Vd) for water at 60°F (velocity, ft/s
0.1
0.2
0.4
0.6 0.8 1
2
4
6
8 10
20
40
×
diameter, in)
60 80 100
200
400 600 800 1000
Values of (Vd) for atmospheric air at 60°F
0.08
4
6 8 10
20
Laminar Critical
flow
zone Transition
zone
0.07
40
60 100
200
400
600 800 1000
2000
4000
8000
6000 10,000
4000
8000
6000 10,000
80,000
20,000 40,000 60,000 100,000
Complete turbulence, rough pipes
0.05
0.04
0.06
Friction factor f =
h
L V2
d 2g
0.04
0.03
flow
inar
Lam 64
f = Re
0.05
0.03
0.02
0.015
0.01
0.008
0.006
Recr
0.004
0.025
0.002
0.001
0.0008
0.0006
0.0004
0.02
Sm
oo
th
0.015
0.0002
pip
es
0.0001
0.000,05
0.01
0.009
0.008
Relative roughness Ɛ
d
0.10
0.09
2
2000
103 2(103) 3
4 56
8104
2(104) 3
4 56
8105
2(105) 3
4 56
8106
2(106) 3
Vd
Reynolds number Re = ν
4 56
8 107
2(107) 3
Ɛ = 0.000,001
d
4 56
0.000,01
8108
Ɛ = 0.000,005
d
Fig. 6.13 The Moody chart for pipe friction with smooth and rough walls. This chart is identical to Eq. (6.48) for turbulent flow
(Courtesy of Richard Collins)
It is accurate to ±15 percent for design calculations over the full range shown in
Fig. 6.13. It can be used for circular and noncircular (Sec. 6.6) pipe flows and
for open-channel flows (Chap. 10). The data can even be adapted as an approximation to boundary layer flows (Chap. 7).
The Moody chart gives a good visual summary of laminar and turbulent pipe
friction, including roughness effects. When the writer was in college, everyone
solved problems by carefully reading this chart. Currently, though, Eq. (6.48),
though implicit in f, is easily solved by iteration or a direct solver. If only a calculator is available, the clever explicit formula given by Haaland [33] as
1
f 1/2
≈ −1.8 log [
6.9
ε/d 1.11
+( ) ]
Red
3.7
(6.49)
varies less than 2 percent from Eq. (6.48).
The shaded area in the Moody chart indicates the range where transition from
laminar to turbulent flow occurs. There are no reliable friction factors in this
range, 2000 < Red, < 4000. Notice that the roughness curves are nearly horizontal in the fully rough regime to the right of the dashed line.
372
Chapter 6 Viscous Flow in Ducts
Table 6.1 Recommended
Roughness Values for Commercial
Ducts
ε
Material
Steel
Iron
Brass
Plastic
Glass
Concrete
Rubber
Wood
Condition
ft
mm
Uncertainty, %
Sheet metal, new
Stainless, new
Commercial, new
Riveted
Rusted
Cast, new
Wrought, new
Galvanized, new
Asphalted cast
Drawn, new
Drawn tubing
—
Smoothed
Rough
Smoothed
Stave
0.00016
0.000007
0.00015
0.01
0.007
0.00085
0.00015
0.0005
0.0004
0.000007
0.000005
Smooth
0.00013
0.007
0.000033
0.0016
0.05
0.002
0.046
3.0
2.0
0.26
0.046
0.15
0.12
0.002
0.0015
Smooth
0.04
2.0
0.01
0.5
±60
±50
±30
±70
±50
±50
±20
±40
±50
±50
±60
±60
±50
±60
±40
From tests with commercial pipes, recommended values for average pipe
roughness are listed in Table 6.1.
EXAMPLE 6.62
Compute the loss of head and pressure drop in 200 ft of horizontal 6-in-diameter
asphalted cast iron pipe carrying water with a mean velocity of 6 ft/s.
Solution
∙
∙
∙
∙
System sketch: See Fig. 6.7 for a horizontal pipe, with Δz = 0 and hf proportional to Δp.
Assumptions: Turbulent flow, asphalted horizontal cast iron pipe, d = 0.5 ft, L = 200 ft.
Approach: Find Red and ε/d; enter the Moody chart, Fig. 6.13; find f, then hf and Δp.
Property values: From Table A.3 for water, converting to BG units, ρ = 998/515.38
= 1.94 slug/ft3, µ = 0.001/47.88 = 2.09 E−5 slug/(ft-s).
∙ Solution step 1: Calculate Red and the roughness ratio. As a crutch, Moody provided
water and air values of “Vd” at the top of Fig. 6.13 to find Red. Instead, let’s calculate it ourselves:
Red =
ρVd (1.94 slug/ft3 ) (6 ft/s) (0.5 ft)
≈ 279,000
=
μ
2.09 E−5 slug/(ft · s)
(turbulent)
From Table 6.1, for asphalted cast iron, ε = 0.0004 ft. Then calculate
ε/d = (0.0004 ft)/(0.5 ft) = 0.0008
∙ Solution step 2: Find the friction factor from the Moody chart or from Eq. (6.48).
If you use the Moody chart, Fig. 6.13, you need practice. Find the line on the right
side for ε/d = 0.0008 and follow it back to the left until it hits the vertical line for
Red ≈ 2.79 E5. Read, approximately, f ≈ 0.02 [or compute f = 0.0198 from Eq. (6.48).]
2
This example was given by Moody in his 1944 paper [8].
6.6 Turbulent Pipe Flow 373
∙ Solution step 3: Calculate hf from Eq. (6.10) and Δp from Eq. (6.8) for a horizontal pipe:
hf = f
(6 ft/s) 2
L V2
200 ft
= (0.02) (
≈ 4.5 ft
d 2g
0.5 ft ) 2(32.2 ft/s2 )
Δp = ρgh f = (1.94 slug/ft3 ) (32.2 ft/s2 ) (4.5 ft) ≈ 280 lbf/ft2
Ans.
Ans.
∙ Comments: In giving this example, Moody [8] stated that this estimate, even for
clean new pipe, can be considered accurate only to about ±10 percent.
EXAMPLE 6.7
Oil, with ρ = 900 kg/m3 and ν = 0.00001 m2/s, flows at 0.2 m3/s through 500 m of
200-mm-diameter cast iron pipe. Determine (a) the head loss and (b) the pressure drop
if the pipe slopes down at 10° in the flow direction.
Solution
First compute the velocity from the known flow rate:
V=
Q
πR
2
=
0.2 m3/s
= 6.4 m/s
π(0.1 m) 2
Then the Reynolds number is
Red =
Vd (6.4 m/s) (0.2 m)
=
= 128,000
v
0.00001 m2/s
From Table 6.1, ε = 0.26 mm for cast iron pipe. Then
ε 0.26 mm
=
= 0.0013
d
200 mm
Enter the Moody chart on the right at ε/d = 0.0013 (you will have to interpolate), and
move to the left to intersect with Re = 128,000. Read f ≈ 0.0225 [from Eq. (6.48) for
these values we could compute f = 0.0227]. Then the head loss is
hf = f
L V2
500 m (6.4 m/s) 2
= (0.0225)
= 117 m
d 2g
0.2 m 2(9.81 m/s2 )
Ans. (a)
From Eq. (6.9) for the inclined pipe,
hf =
Δp
Δp
+ z1 − z2 =
+ L sin 10°
ρg
ρg
or Δp = ρg[hf − (500 m) sin 10°] = ρg(117 m − 87 m)
= (900 kg/m3 ) (9.81 m/s2 ) (30 m) = 265,000 kg/(m · s2 ) = 265,000 Pa Ans. (b)
EXAMPLE 6.8
Repeat Example 6.4 to see whether there is any possible turbulent flow solution for a
smooth-walled pipe.
374
Chapter 6 Viscous Flow in Ducts
Solution
In Example 6.4 we estimated a head loss hf ≈ 1.66 ft, assuming laminar exit flow (α ≈
2.0). For this condition the friction factor is
f = hf
(0.004 ft) (2) (32.2 ft/s2 )
d 2g
=
(1.66
ft)
≈ 0.0388
L V2
(1.0 ft) (3.32 ft/s) 2
For laminar flow, Red = 64/f = 64/0.0388 ≈ 1650, as we showed in Example 6.4. However,
from the Moody chart (Fig. 6.13), we see that f = 0.0388 also corresponds to a turbulent smooth-wall condition, at Red ≈ 4500. If the flow actually were turbulent, we should
change our kinetic energy factor to α ≈ 1.06 [Eq. (3.77)], whence the corrected hf ≈ 1.82 ft
and f ≈ 0.0425. With f known, we can estimate the Reynolds number from our formulas:
Red ≈ 3250
[Eq. (6.38) ]
or
Red ≈ 3400
[Eq. (6.39 b) ]
So the flow might have been turbulent, in which case the viscosity of the fluid would have been
μ=
ρVd 1.80(3.32) (0.004)
=
= 7.2 × 10−6 slug/(fts )
Red
3300
Ans.
This is about 55 percent less than our laminar estimate in Example 6.4. The moral is to keep
the capillary-flow Reynolds number below about 1000 to avoid such duplicate solutions.
6.7 Four Types of Pipe Flow Problems
The head loss equation (6.10) and the Moody chart in Fig. 6.13 (or the Colebrook
equation (6.48)) can be used to solve almost any problem involving friction losses
in long pipe flows. In all instances, we assume that the fluid properties ρ, µ, and
surface roughness ε are given. Equation (6.10) and the Moody chart together can
solve two unknowns. One is the friction factor f, and the other is one of four
fundamental problems that are commonly encountered in pipe flow calculations:
1. Given d, L, and V or Q, ρ, µ, and g, compute the head loss hf (head loss
problem).
2. Given d, L, hf, ρ, µ, and g, compute the velocity V or flow rate Q (flow
rate problem).
3. Given Q, L, hf, ρ, µ, and g, compute the diameter d of the pipe (sizing problem).
4. Given Q, d, hf, ρ, µ, and g, compute the pipe length L.
Problems of types 1 and 4 are straightforward. They are well suited to the Moody
chart. However, problems of types 2 and 3 involve iteration and repeated calculations using the chart. In the problem of the second type, for example, the diameter is given but the flow rate is unknown. An initial guess for the friction factor
in that case is from the turbulent flow region for the given roughness. Once the
velocity (flow rate) is obtained from Eq. (6.10), the friction factor is corrected
using the Moody chart, and the process is repeated until the solution converges.
There are two alternatives to iteration for problems of types 2 and 3: (a)
preparation of a suitable new Moody-type formula (see Probs. P6.68 and P6.73);
or (b) the use of solver software, like Excel. Examples 6.9 and 6.11 include the
Excel approach to these problems.
6.7 Four Types of Pipe Flow Problems 375
Type 2 Problem: Find the Flow Rate
Even though velocity (or flow rate) appears in both the ordinate and the abscissa
on the Moody chart, iteration for turbulent flow is nevertheless quite fast because
f varies so slowly with Red. In earlier editions, the writer rescaled the Colebrook
formula (6.48) into a relation where Q could be calculated directly. That idea is
now downsized to Prob. P6.68. Example 6.9, which follows, is illustrated both
by iteration and by an Excel solution.
EXAMPLE 6.9
Oil, with ρ = 950 kg/m3 and ν = 2 E-5 m2/s, flows through a 30-cm-diameter pipe
100 m long with a head loss of 8 m. The roughness ratio is ε/d = 0.0002. Find the
average velocity and flow rate.
Iterative Solution
By definition, the friction factor is known except for V:
f = hf
2(9.81 m/s2 )
d 2g
0.3 m
=
(8
m)
( 100 m )[
] or
L V2
V2
f V2 ≈ 0.471
(SI units)
To get started, we only need to guess f, compute V = √0.471/f , then get Red, compute a
better f from the Moody chart, and repeat. The process converges fairly rapidly. A good
first guess is the “fully rough” value for ε/d = 0.0002, or f ≈ 0.014 from Fig. 6.13. The
iteration would be as follows:
Guess f ≈ 0.014, then V = √0.471/0.014 = 5.80 m/s and Red = Vd/ν ≈ 87,000.
At Red = 87,000 and ε/d = 0.0002, compute fnew ≈ 0.0195 [Eq. (6.48)].
New f ≈ 0.0195, V = √0.471/0.0195 = 4.91 m/s and Red = Vd/ν = 73,700. At
Red = 73,700 and ε/d = 0.0002, compute fnew ≈ 0.0201 [Eq. (6.48)].
Better f ≈ 0.0201, V = √0.471/0.0201 = 4.84 m/s and Red ≈ 72,600. At
Red = 72,600 and ε/d = 0.0002, compute fnew ≈ 0.0201 [Eq. (6.48)].
We have converged to three significant figures. Thus our iterative solution is
V = 4.84 m/s
π
π
Q = V ( ) d2 = (4.84) ( ) (0.3) 2 ≈ 0.342 m3/s
4
4
Ans.
The iterative approach is straightforward and not too onerous, so it is routinely used by
e­ ngineers. Obviously this repetitive procedure is ideal for a personal computer.
Solution by Iteration with Excel
To iterate by repeated copying in Excel, we need five columns: velocity, flow rate,
­Reynolds number, an initial guess for f, and a calculation of f from (ε/d) = 0.0002 and
the current value of Red. We modify our guess for f, in the next row, with the new value of
f and calculate again, as shown in the following table. Since f is a slowly varying function,
the process converges rapidly.
376
Chapter 6 Viscous Flow in Ducts
V(m/s) =
(0.471/E1)0.5
Q(m3/s) =
(π/4)A1*0.32
Red =
A1*0.3/0.00002
f(Eq. 6.48)
f-guess
A
B
C
D
E
1
2
3
5.8002
4.8397
4.8397
0.4100
0.3421
0.3421
87004
72596
72596
0.02011
0.02011
0.02011
0.01400
0.02011
0.02011
As shown in the hand-iterated method, the proper solution is V = 4.84 m/s and
Q = 0.342 m2/s.
Type 3 Problem: Find the Pipe Diameter
The Moody chart is especially awkward for finding the pipe size, since d occurs
in all three parameters f, Red, and ε/d. Further, it depends on whether we know
the velocity or the flow rate. We cannot know both, or else we could immediately
compute d = √4Q/(πV).
Let us assume that we know the flow rate Q. Note that this requires us to
redefine the Reynolds number in terms of Q:
Red =
4Q
Vd
=
ν
πdν
(6.50)
If, instead, we knew the velocity V, we could use the first form for the Reynolds
number. The writer finds it convenient to solve the Darcy friction factor correlation, Eq. (6.10), by solving for f :
f = hf
5
d 2g π2 ghf d
=
L V2
8 LQ2
(6.51)
The following two examples illustrate the iteration.
EXAMPLE 6.10
Work Example 6.9 backward, assuming that Q = 0.342 m3/s and ε = 0.06 mm are
known but that d (30 cm) is unknown. Recall L = 100 m, ρ = 950 kg/m3, ν = 2 E−5
m2/s, and hf = 8 m.
Iterative Solution
First write the diameter in terms of the friction factor:
f=
π 2 (9.81 m/s2 ) (8 m)d 5
= 8.28d 5 or
8 (100 m) (0.342 m3/s) 2
d ≈ 0.655f 1/5
(1)
6.7 Four Types of Pipe Flow Problems 377
in SI units. Also write the Reynolds number and roughness ratio in terms of the diameter:
Red =
4(0.342 m3/s)
=
π(2 E−5 m2/s)d
ε 6 E−5 m
=
d
d
21,800
d
(2)
(3)
Guess f, compute d from (1), then compute Red from (2) and ε/d from (3), and compute
a better f from the Moody chart or Eq. (6.48). Repeat until (fairly rapid) convergence.
Having no initial estimate for f, the writer guesses f ≈ 0.03 (about in the middle of the
turbulent portion of the Moody chart). The following calculations result:
f ≈ 0.03
Red ≈
Eq. (6.48):
d ≈ 0.655(0.03) 1/5 ≈ 0.325 m
21,800
≈ 67,000
0.325
fnew ≈ 0.0203
then
fbetter ≈ 0.0201
dnew ≈ 0.301 m
ε
≈ 2.0 E−4
d
Red,new ≈ 72,500
Eq. (6.48):
ε
≈ 1.85 E−4
d
and
d = 0.300 m
Ans.
The procedure has converged to the correct diameter of 30 cm given in Example 6.9.
Solution by Iteration with Excel
To iterate by repeated copying in Excel, we need five columns: ε/d, friction factor, Reynolds number, diameter d, and an initial guess for f. With the guess for f, we calculate d ≈
0.655f 1/5, Red ≈ 21,800/d, and ε/d = (0.00006 m)/d. Replace the guessed f with the new
f. Thus Excel is doing the work of our previous hand calculation:
ε/d = 0.00006/d
f − Eq. (6.48)
A
1
0.000185
2
0.000201
3
0.000200
4
0.000200
Red =
21,800/d
d(meters) =
0.655f 0.2
f-guess
B
C
D
E
0.0196
0.0201
0.0201
0.0201
67111
73106
72677
72706
0.325
0.298
0.300
0.300
0.0300
0.0196
0.0201
0.0201
As shown in our hand-iterated method, the proper solution is d = 0.300 m.
EXAMPLE 6.11
A smooth plastic pipe is to be designed to carry 8 ft3/s of water at 20°C through 1000
ft of horizontal pipe with an exit at 15 lbf/in2. The pressure drop is to be approximately
250 lbf/in2. Determine (a) the proper diameter for this pipe and (b) whether a Schedule
40 is suitable if the pipe material has an allowable stress of 8000 lbf/in2.
378
Chapter 6 Viscous Flow in Ducts
Solution by Excel Iteration
Assumptions: Steady turbulent flow, smooth walls. For water, take ρ = 1.94 slug/ft3 and
µ = 2.09 E−5 slug/(ft · s). With d unknown, use Eq. (6.51):
f=
5
5
(250 × 144 lbf/ft2 )d 5
π 2 ghf d
π2 Δpd
π2
=
=
= 0.358 d 5 (1)
2
2
8 LQ
8 ρLQ
8 (1.94 slug/ft3 ) (1000 ft) (8 ft3/s) 2
We know neither d nor f, but they are related by the Prandtl formula, Eq. (6.38):
1
f
1/2
≈ 2.0 log(Red f 1/2 ) − 0.8, Red =
ρVd 4ρQ
4(1.94) (8)
945,500
=
=
(2)
=
μ
π μ d π (2.09 E−5)d
d
Part (a)
Equations (1) and (2) can be solved simultaneously for f and d. Using Excel iteration, we
have four columns: a guessed f = 0.02, d from Eq. (1), Red from Eq. (2), and a better f
from Eq. (6.38). The pipe is smooth, so we don’t need roughness:
f − Eq. (6.38)
Red = 945500/C2
d = (D2/0.358)0.2
f-guess
A
B
C
D
1
2
3
4
0.01009
0.01047
0.01044
0.01045
1683574
1930316
1916418
1917156
0.562
0.490
0.493
0.493
0.02000
0.01009
0.01047
0.01044
The process converges rapidly to
Red ≈ 1.92 E6;
f ≈ 0.01045;
d ≈ 0.493 ft
Take the next highest Schedule 40 diameter in Table 6.2: d ≈ 0.5 ft = 6 in
Ans. (a)
Part (b)
Check to see if Schedule 40 is strong enough. The maximum pressure occurs at the pipe
­entrance: pmax = pexit + Δp = 15 + 250 = 265 lb/in2. The schedule number is thus
Schedule number = (1000)
265 psi
(maximum pressure)
= (1000) (
≈ 33
(allowable stress)
8000 psi )
Schedule 30 is too weak for this pressure, so choose a Schedule 40 pipe.
Commercial Pipe Sizes
Ans. (b)
In solving a problem to find the pipe diameter, we should note that commercial
pipes are made only in certain sizes. Table 6.2 gives nominal and actual sizes of
pipes in the United States. The term Schedule 40 is a measure of the pipe thickness and its resistance to stress caused by internal fluid pressure. If P is the
internal fluid pressure and S is the allowable stress of the pipe material, then the
6.7 Four Types of Pipe Flow Problems 379
Table 6.2 Nominal and Actual
Sizes of a Schedule 40 Pipe
Nominal size, in
Actual ID, in
1/80.269
1/40.364
3/80.493
1/20.622
3/40.824
11.049
1-1/21.610
22.067
2-1/22.469
33.068
44.026
55.047
66.065
Wall thickness, in
0.068
0.088
0.091
0.109
0.113
0.133
0.145
0.154
0.203
0.216
0.237
0.258
0.280
schedule number = (1000)(P/S). Commercial schedules vary from 5 to 160, but
40 and 80 are by far the most popular. Example 6.11 is a typical application.
Type 4 Problem: Find the Pipe Length
In designing piping systems, it is desirable to estimate the appropriate pipe length
for a given pipe diameter, pump power, and flow rate. The pump head will match
the piping head loss. If minor losses, Sec. 6.9, are neglected, the (horizontal) pipe
length follows from Darcy’s formula (6.10):
Power
L V2
= hf = f
(6.52)
ρgQ
d 2g
With Q, d, and ε known, we may compute Red and f, after which L is obtained
from the formula. Note that pump efficiency varies strongly with flow rate
(Chap. 11). Thus, it is important to match pipe length to the pump’s region of
maximum ­efficiency.
hpump =
EXAMPLE 6.12
A pump delivers 0.6 hp to water at 68°F, flowing in a 6-in-diameter asphalted cast iron
horizontal pipe at V = 6 ft/s. What is the proper pipe length to match these conditions?
Solution
∙ Approach: Find hf from the known power and find f from Red and ε/d. Then find L.
∙ Water properties: For water at 68°F, Table A.3, converting to BG units, ρ = 1.94 slug/
ft3 and µ = 2.09 E−5 slug/(ft · s).
∙ Pipe roughness: From Table 6.1 for asphalted cast iron, ε = 0.0004 ft.
∙ Solution step 1: Find the pump head from the flow rate and the pump power:
π
ft
ft3
(0.5 ft) 2 (6 ) = 1.18
s
s
4
(0.6 hp) [550(ft · lbf )/(s · hp) ]
Q = AV =
h pump =
Power
=
= 4.48 ft
ρgQ
(1.94 slug/ft3 ) (32.2 ft/s2 ) (1.18 ft3/s)
380
Chapter 6 Viscous Flow in Ducts
∙ Solution step 2:
Red =
Compute the friction factor from the Colebrook formula, Eq. (6.48):
ρVd (1.94) (6) (0.5)
=
= 278,500
μ
2.09 E−5
1
√f
≈ −2.0 log10 (
ε 0.0004 ft
=
= 0.0008
d
0.5 ft
ε/d
2.51
+
yields
3.7 Red √f )
f = 0.0198
∙ Solution step 3: Find the pipe length from the Darcy formula (6.10):
hp = hf = 4.48 ft = f
(6 ft/s) 2
L V2
L
= (0.0198) (
)
d 2g
0.5 ft 2(32.2 ft/s2 )
Solve for
L ≈ 203 ft
Ans.
∙ Comment: This is Moody’s problem (Example 6.6) turned around so that the length is
unknown.
6.8 Flow in Noncircular Ducts3
If the duct is noncircular, the analysis of fully developed flow follows that of the
circular pipe but is more complicated algebraically. For laminar flow, one can
solve the exact equations of continuity and momentum. For turbulent flow, the
logarithm law velocity profile can be used, or (better and simpler) the hydraulic
diameter is an excellent approximation.
The Hydraulic Diameter
For a noncircular duct, the control volume concept of Fig. 6.7 is still valid, but
the cross-sectional area A does not equal πR2 and the cross-sectional perimeter
wetted by the shear stress 3 does not equal 2πR. The momentum equation (6.9a)
thus becomes
Δp A + ρgA ΔL sin ϕ − τw3ΔL = 0
or
hf =
Δp
τw ΔL
+ Δz =
ρg
ρg A/3
(6.53)
Comparing this to Eq. (6.9b), we see that A/ 3 takes the place of one-fourth of
the pipe diameter for a circular cross section. We define the friction factor in
terms of average shear stress:
fNCD =
8τw
ρV2
(6.54)
where NCD stands for noncircular duct and V = Q/A as usual, Eq. (6.53) becomes
hf = f
3
L V2
Dh 2g
This section may be omitted without loss of continuity.
(6.55)
6.8 Flow in Noncircular Ducts 381
This is equivalent to Eq. (6.10) for pipe flow except that d is replaced by Dh.
Therefore, we customarily define the hydraulic diameter as
4A
4 × area
=
3
wetted perimeter
Dh =
(6.56)
We should stress that the wetted perimeter includes all surfaces acted upon by
the shear stress. For example, in a circular annulus, both the outer and the inner
perimeters should be added.
We would therefore expect by dimensional analysis that this friction factor f,
based on hydraulic diameter as in Eq. (6.55), would correlate with the Reynolds
number and roughness ratio based on the hydraulic diameter
VDh ε
,
f=F(
ν Dh )
(6.57)
and this is the way the data are correlated. But we should not necessarily expect
the Moody chart (Fig. 6.13) to hold exactly in terms of this new length scale.
And it does not, but it is surprisingly accurate:
f≈ µ
64
ReDh
fMoody (ReDh,
ε
Dh )
±40%
laminar flow
±15%
turbulent flow
(6.58)
Now let us look at some particular cases.
Flow between Parallel Plates
Probably the simplest noncircular duct flow is fully developed flow between
parallel plates a distance 2h apart, as in Fig. 6.14. As noted in the figure, the
width b ≫ h, so the flow is essentially two-dimensional; that is, u = u(y) only.
The hydraulic diameter is
Dh =
4(2bh)
4A
= lim
= 4h
b→∞ 2b + 4h
3
b ⃪
y = +h
y
2h
u ( y)
x
Y
Fig. 6.14 Fully developed flow
­between parallel plates.
u max
y=–h
∞
(6.59)
382
Chapter 6 Viscous Flow in Ducts
that is, twice the distance between the plates. The pressure gradient is constant,
(−dp/dx) = Δp/L, where L is the length of the channel along the x axis.
Laminar Flow Solution
The laminar solution was given in Sec. 4.10, in connection with Fig. 4.16b. Let
us review those results here:
u = umax (1 −
h2 Δp
where
u
=
max
2μ L
h2 )
y2
2bh3 Δp
3μ L
Q h2 Δp 2
V= =
= umax
A 3μ L
3
Δp 3μV
du
τw = μ ` `
=h
=
dy y=h
L
h
Δp 3μLV
=
hf =
ρg
ρgh2
Q=
Now use the head loss to establish the laminar friction factor:
hf
96μ
96
flam =
=
=
2
(L/Dh )(V /2g) ρV(4h) ReDh
(6.60)
(6.61)
Thus, if we could not work out the laminar theory and chose to use the approximation f ≈ 64/ReDh, we would be 33 percent low. The hydraulic-diameter approximation is relatively crude in laminar flow, as Eq. (6.58) states.
Just as in circular-pipe flow, the laminar solution above becomes unstable at
about ReDh ≈ 2000; transition occurs and turbulent flow results.
Turbulent Flow Solution
For turbulent flow between parallel plates, we can again use the logarithm law,
Eq. (6.28), as an approximation across the entire channel, using not y but a wall
coordinate Y, as shown in Fig. 6.14:
u(Y) 1 Yu*
≈ ln
+B
0 < Y < h
(6.62)
κ
ν
u*
This distribution looks very much like the flat turbulent profile for pipe flow in
Fig. 6.11b, and the mean velocity is
V=
1
h
∫ u dY = u*(1κ ln hu*ν + B − 1κ )
h
0
(6.63)
Recalling that V/u* = (8/f)1/2, we see that Eq. (6.63) is equivalent to a parallelplate friction law. Rearranging and cleaning up the constant terms, we obtain
1
≈ 2.0 log (ReDh f 1/2 ) − 1.19
(6.64)
f 1/2
6.8 Flow in Noncircular Ducts 383
where we have introduced the hydraulic diameter Dh = 4h. This is remarkably
close to the smooth-wall pipe friction law, Eq. (6.38). Therefore we conclude that
the use of the hydraulic diameter in this turbulent case is quite successful. That
turns out to be true for other noncircular turbulent flows also.
Equation (6.64) can be brought into exact agreement with the pipe law by
rewriting it in the form
1
f 1/2
= 2.0 log (0.64 ReDh f 1/2 ) − 0.8
(6.65)
Thus the turbulent friction is predicted most accurately when we use an effective
diameter Deff equal to 0.64 times the hydraulic diameter. The effect on f itself is
much less, about 10 percent at most. We can compare with Eq. (6.66) for laminar
flow, which predicted
Parallel plates:
Deff =
64
2
Dh = Dh 96
3
(6.66)
This close resemblance (0.64Dh versus 0.667Dh) occurs so often in noncircular duct flow that we take it to be a general rule for computing turbulent
friction in ducts:
4A
reasonable accuracy
3
64
Deff = Dh
better accuracy
( f ReDh )laminar theory
Deff = Dh =
(6.67)
Jones [10] shows that the effective-laminar-diameter idea collapses all data for
rectangular ducts of arbitrary height-to-width ratio onto the Moody chart for pipe
flow. We recommend this idea for all noncircular ducts.
EXAMPLE 6.13
Fluid flows at an average velocity of 6 ft/s between horizontal parallel plates a distance
of 2.4 in apart. Find the head loss and pressure drop for each 100 ft of length for ρ =
1.9 slugs/ft3 and (a) ν = 0.00002 ft2/s and (b) ν = 0.002 ft2/s. Assume smooth walls.
Solution
Part (a)
The viscosity µ = ρν = 3.8 × 10−5 slug/(ft · s). The spacing is 2h = 2.4 in = 0.2 ft, and Dh
= 4h = 0.4 ft. The Reynolds number is
ReDh =
VDh (6.0 ft/s) (0.4 ft)
=
= 120,000
ν
0.00002 ft2/s
The flow is therefore turbulent. For reasonable accuracy, simply look on the Moody
chart (Fig. 6.13) for smooth walls:
f ≈ 0.0173 hf ≈ f
L V2
100 (6.0) 2
= 0.0173
≈ 2.42 ft
Dh 2g
0.4 2(32.2)
Ans. (a)
384
Chapter 6 Viscous Flow in Ducts
Since there is no change in elevation,
Δp = ρghf = 1.9(32.2) (2.42) = 148 lbf/ft2
Ans. (a)
This is the head loss and pressure drop per 100 ft of channel. For more accuracy, take
Deff = 64
96 Dh from laminar theory; then
Reeff = 64
96 (120,000) = 80,000
and from the Moody chart read f ≈ 0.0189 for smooth walls. Thus a better estimate is
hf = 0.0189
100 (6.0) 2
= 2.64 ft
0.4 2(32.2)
Δp = 1.9(32.2) (2.64) = 161 lbf/ft2
and
Better ans. (a)
The more accurate formula predicts friction about 9 percent higher.
Part (b)
Compute µ = ρν = 0.0038 slug/(ft · s). The Reynolds number is 6.0(0.4)/0.002 = 1200;
therefore the flow is laminar, since Re is less than 2300.
You could use the laminar flow friction factor, Eq. (6.61)
flam =
from which
96
96
=
= 0.08
ReDh 1200
hf = 0.08
100 (6.0) 2
= 11.2 ft
0.4 2(32.2)
Δp = 1.9(32.2) (11.2) = 684 lbf/ft2
and
Ans. (b)
Alternately you can finesse the Reynolds number and go directly to the appropriate
laminar flow formula, Eq. (6.60):
V=
or
Δp =
h2 Δp
3μ L
3(6.0 ft/s) [0.0038 slug/(ft · s) ] (100 ft)
and
(0.1 ft) 2
hf =
= 684 slugs/(ft · s2 ) = 684 lbf/ft2
Δp
684
=
= 11.2 ft
ρg
1.9(32.2)
Flow through a Concentric Annulus
Consider steady axial laminar flow in the annular space between two concentric
­cylinders, as in Fig. 6.15. There is no slip at the inner (r = b) and outer radii (r = a).
For u = u(r) only, the governing relation is Eq. (D.7) in Appendix D:
d
du
d
rμ ) = Kr K =
(p + ρgz)
(
dr
dr
dx
Integrate this twice:
u=
1 2K
+ C1 ln r + C2
r
4 μ
(6.68)
6.8 Flow in Noncircular Ducts 385
r=a
r
u(r)
r=b
x
u(r)
Fig. 6.15 Fully developed flow
through a concentric annulus.
The constants are found from the two no-slip conditions:
1
u(r = a) = 0 = a2
4
1
u(r = b) = 0 = b2
4
K
+ C1 ln a + C2
μ
K
+ C1 ln b + C2
μ
The final solution for the velocity profile is
u=
1
d
a2 − b2 a
2
2
−
(p
+
ρgz)
a
−
r
+
ln (6.69)
][
4μ [ dx
ln (b/a) r ]
The volume flow is given by
Q=
∫
a
u 2πr dr =
b
(a2 − b2 ) 2
π
d
4
4
−
(p
+
ρgz)
a
−
b
−
(6.70)
][
8μ [ dx
ln (a/b) ]
The velocity profile u(r) resembles a parabola wrapped around in a circle to form
a split doughnut, as in Fig. 6.15.
It is confusing to base the friction factor on the wall shear because there are
two shear stresses, the inner stress being greater than the outer. It is better to
define f with respect to the head loss, as in Eq. (6.55),
f = hf
Dh 2g
L V2
where V =
Q
2
π(a − b2 )
(6.71)
The hydraulic diameter for an annulus is
Dh =
4π(a2 − b2 )
= 2(a − b)(6.72)
2π(a + b)
It is twice the clearance, rather like the parallel-plate result of twice the distance
between plates [Eq. (6.59)].
Substituting hf, Dh, and V into Eq. (6.71), we find that the friction factor for
laminar flow in a concentric annulus is of the form
f=
64ζ
ReDh
ζ=
(a − b) 2 (a2 − b2 )
a4 − b4 − (a2 − b2 ) 2/ln (a/b)
(6.73)
386
Chapter 6 Viscous Flow in Ducts
Table 6.3 Laminar Friction Factors
for a Concentric Annulus
b/a
0.0
0.00001
0.0001
0.001
0.01
0.05
0.1
0.2
0.4
0.6
0.8
1.0
f ReDh
Deff/Dh = 1/ζ
64.0
70.09
71.78
74.68
80.11
86.27
89.37
92.35
94.71
95.59
95.92
96.0
1.000
0.913
0.892
0.857
0.799
0.742
0.716
0.693
0.676
0.670
0.667
0.667
The dimensionless term ζ is a sort of correction factor for the hydraulic diameter.
We could rewrite Eq. (6.73) as
Concentric annulus:
f=
64
Reeff
Reeff =
1
ReDh(6.74)
ζ
Some numerical values of f ReDh and Deff/Dh = 1/ζ are given in Table 6.3. Again,
laminar annular flow becomes unstable at ReDh ≈ 2000.
For turbulent flow through a concentric annulus, the analysis might proceed
by patching together two logarithmic law profiles, one going out from the inner
wall to meet the other coming in from the outer wall. We omit such a scheme
here and proceed directly to the friction factor. According to the general rule
proposed in Eq. (6.58), turbulent friction is predicted with excellent accuracy
by replacing d in the Moody chart with Deff = 2(a − b)/ζ, with values listed
in Table 6.3.4 This idea also includes roughness (replace ε/d in the chart with
ε/Deff). For a quick design number with about 10 percent accuracy, one can simply use the hydraulic diameter Dh = 2(a − b).
EXAMPLE 6.14
What should the reservoir level h be to maintain a flow of 0.01 m3/s through the commercial steel annulus 30 m long shown in Fig. E6.14? Neglect entrance effects and
take ρ = 1000 kg/m3 and ν = 1.02 × 10−6 m2/s for water.
1
h=?
a = 5 cm
b = 3 cm
2
Q, V
E6.14
Water
L = 30 m
Solution
∙
∙
∙
∙
Assumptions: Fully developed annulus flow, minor losses neglected.
Approach: Determine the Reynolds number, then find f and hf and thence h.
Property values: Given ρ = 1000 kg/m3 and ν = 1.02 E−6 m2/s.
Solution step 1: Calculate the velocity, hydraulic diameter, and Reynolds number:
Q
0.01 m3/s
m
=
= 1.99
s
A π[ (0.05 m) 2 − (0.03 m) 2 ]
Dh = 2(a − b) = 2(0.05 m − 0.03 m) = 0.04 m
VDh (1.99 m/s) (0.04 m)
ReDh =
=
= 78,000 (turbulent flow)
ν
1.02 E-6 m2/s
V=
4
Jones and Leung [44] show that data for annular flow also satisfy the effective-laminar-­
diameter idea.
6.8 Flow in Noncircular Ducts 387
∙ Solution step 2: Apply the steady flow energy equation between sections 1 and 2:
p1 α1V21
p2 α2V22
+
+ z1 =
+
+ z2 + hf
ρg
ρg
2g
2g
or
h=
α2V22
V22
L
+ hf =
α2 + f ) (
2g
2g
Dh
(1)
Note that z1 = h. For turbulent flow, from Eq. (3.43c), we estimate α2 ≈ 1.03
∙ Solution step 3: Determine the roughness ratio and the friction factor. From Table
6.1, for (new) commercial steel pipe, ε = 0.046 mm. Then
ε
0.046 mm
=
= 0.00115
Dh
40 mm
For a reasonable estimate, use ReDh to estimate the friction factor from Eq. (6.48):
1
√f
≈ −2.0 log10 (
0.00115
2.51
+
solve for f ≈ 0.0232
3.7
78,000 √f )
For slightly better accuracy, we could use Deff = Dh /ζ. From Table 6.3, for b/a = 3/5,
1/ζ = 0.67. Then Deff = 0.67(40 mm) = 26.8 mm, whence ReDeff = 52,300, ε/Deff =
0.00172, and feff ≈ 0.0257. Using the latter estimate, we find the required reservoir level
from Eq. (1):
h=
V22
(1.99 m/s) 2
L
30 m
α
+
f
=
1.03 + 0.0257
≈ 4.1 m
2
eff
2g (
Dh ) 2(9.81 m/s) 2 [
0.04 m ]
Ans.
∙ Comments: Note that we do not replace Dh with Deff in the head loss term fL/Dh,
which comes from a momentum balance and requires hydraulic diameter. If we used
the simpler friction estimate, f ≈ 0.0232, we would obtain h ≈ 3.72 m, or about
9 percent lower.
Table 6.4 Laminar Friction
Constants f Re for Rectangular and
Triangular Ducts
Rectangular
b
Isosceles triangle
2θ
a
b/a
f ReDh
0.0
0.05
0.1
0.125
0.167
0.25
0.4
0.5
0.75
1.0
96.00 0
89.91
10
84.68
20
82.34
30
78.81
40
72.93
50
65.47
60
62.19
70
57.89
80
56.91
90
θ, deg
f ReDh
48.0
51.6
52.9
53.3
52.9
52.0
51.1
49.5
48.3
48.0
Other Noncircular Cross Sections
In principle, any duct cross section can be solved analytically for the laminar flow
velocity distribution, volume flow, and friction factor. This is because any cross
section can be mapped onto a circle by the methods of complex variables, and
other powerful analytical techniques are also available. Many examples are given
by White [3, pp. 112–115], Berker [11], and Olson [12]. Reference 34 is devoted
entirely to laminar duct flow.
In general, however, most unusual duct sections have strictly academic and not
commercial value. We list here only the rectangular and isosceles-triangular sections, in Table 6.4, leaving other cross sections for you to find in the references.
For turbulent flow in a duct of unusual cross section, one should replace d with
Dh on the Moody chart if no laminar theory is available. If laminar results are
known, such as Table 6.4, replace d with Deff = [64/(f Re)]Dh for the particular
geometry of the duct.
For laminar flow in rectangles and triangles, the wall friction varies greatly,
being largest near the midpoints of the sides and zero in the corners. In turbulent
388
Chapter 6 Viscous Flow in Ducts
Midplane
(a)
(b)
Fig. 6.16 Illustration of secondary turbulent flow in noncircular ducts:
(a) axial mean velocity contours; (b) secondary flow in-plane cellular
motions. (After J. Nikuradse, dissertation, Gottingen, 1926.)
flow through the same sections, the shear is nearly constant along the sides, dropping off sharply to zero in the corners. This is because of the phenomenon of
turbulent secondary flow, in which there are nonzero mean velocities v and w in
the plane of the cross section. Some measurements of axial velocity and secondary
flow patterns are shown in Fig. 6.16, as sketched by Nikuradse in his 1926 dissertation. The secondary flow “cells” drive the mean flow toward the corners, so
that the axial velocity contours are similar to the cross section and the wall shear
is nearly constant. This is why the hydraulic-diameter concept is so successful for
turbulent flow. Laminar flow in a straight noncircular duct has no secondary flow.
An accurate theoretical prediction of turbulent secondary flow has yet to be
achieved, although numerical models are often successful [36].
EXAMPLE 6.15
Air, with ρ = 0.00237 slug/ft3 and ν = 0.000157 ft2/s, is forced through a horizontal
square 9-by-9-in duct 100 ft long at 25 ft3/s. Find the pressure drop if ε = 0.0003 ft.
Solution
Compute the mean velocity and hydraulic diameter:
25 ft3/s
= 44.4 ft/s
(0.75 ft) 2
4A 4(81 in2 )
Dh =
=
= 9 in = 0.75 ft
3
36 in
V=
6.9 Minor or Local Losses in Pipe Systems 389
From Table 6.4, for b/a = 1.0, the effective diameter is
64
Deff =
Dh = 0.843 ft
56.91
VDeff 44.4(0.843)
whence
Reeff =
=
= 239,000
ν
0.000157
ε
0.0003
=
= 0.000356
Deff
0.843
From the Moody chart, read f = 0.0177. Then the pressure drop is
Δp = ρghf = ρg (f
or
L V2
100 44.42
=
0.00237(32.2)
0.0177
[
Dh 2g )
0.75 2(32.2) ]
Δp = 5.5 lbf/ft2
Ans.
Pressure drop in air ducts is usually small because of the low density.
6.9 Minor or Local Losses in Pipe Systems
For any pipe system, in addition to the Moody-type friction loss computed for the
length of pipe, there are additional so-called minor losses or local losses due to
1.
2.
3.
4.
5.
Pipe entrance or exit.
Sudden expansion or contraction.
Bends, elbows, tees, and other fittings.
Valves, open or partially closed.
Gradual expansions or contractions.
The losses may not be so minor; for example, a partially closed valve can cause
a greater pressure drop than a long pipe.
Since the flow pattern in fittings and valves is quite complex, the theory is
very weak. The losses are commonly measured experimentally and correlated
with the pipe flow parameters. The data, especially for valves, are somewhat
dependent on the particular manufacturer’s design, so that the values listed here
must be taken as average design estimates [15, 16, 35, 43, 46].
The measured minor loss is usually given as a ratio of the head loss hm = Δp/
(ρg) through the device to the velocity head V2/(2g) of the associated piping system:
Δp
hm
Loss coefficient K = 2
= 1 2 (6.75)
V /(2g) 2ρV
Although K is dimensionless, it often is not correlated in the literature with the
Reynolds number and roughness ratio but rather simply with the raw size of the
pipe in, say, inches. Almost all data are reported for turbulent flow conditions.
A single pipe system may have many minor losses. Since all are correlated
with V2/(2g), they can be summed into a single total system loss if the pipe has
constant diameter:
Δhtot = hf + Σhm =
V2 fL
+ ΣK) (6.76)
2g ( d
390
Chapter 6 Viscous Flow in Ducts
h
h
D
(a)
·
D
D
(b)
D
D
(d)
Fig. 6.17 Typical commercial valve
geometries: (a) gate valve; (b) globe
valve; (c) angle valve; (d) swingcheck valve; (e) disk-type gate
valve.
h
D
D
(c)
(e)
Note, however, that we must sum the losses separately if the pipe size changes
so that V2 changes. The length L in Eq. (6.76) is the total length of the pipe axis.
There are many different valve designs in commercial use. Figure 6.17 shows five
typical designs: (a) the gate, which slides down across the section; (b) the globe,
which closes a hole in a special insert; (c) the angle, similar to a globe but with a
90° turn; (d) the swing-check valve, which allows only one-way flow; and (e) the
disk, which closes the section with a circular gate. The globe, with its tortuous flow
path, has the highest losses when fully open. Many excellent details about these and
other valves are given in the handbooks by Skousen [35] and Crane Co. [52].
Table 6.5 lists loss coefficients K for four types of valve, three angles of elbow
fitting, and two tee connections. Fittings may be connected by either internal
screws or flanges, hence the two listings. We see that K generally decreases
with pipe size, which is consistent with the higher Reynolds number and decreased
roughness ratio of large pipes. We stress that Table 6.5 represents losses ­averaged
among various manufacturers, so there is an uncertainty as high as ±50 percent.
In addition, most of the data in Table 6.5 are relatively old [15, 16] and therefore
based on fittings manufactured in the 1950s. Modern forged and molded fittings
may yield somewhat different loss factors, often less than those listed in Table 6.5.
An example, shown in Fig. 6.18a, gives recent data [48] for fairly short (bendradius/elbow-diameter = 1.2) flanged 90° elbows. The elbow diameter was 1.69 in.
Notice first that K is plotted versus Reynolds number, rather than versus the raw
(dimensional) pipe diameters in Table 6.5, and therefore Fig. 6.18a has more
generality. Then notice that the K values of 0.23 ± 0.05 are significantly less than
6.9 Minor or Local Losses in Pipe Systems 391
Table 6.5 Resistance Coefficients
K = hm/[V2/(2g)] for Open Valves,
Elbows, and Tees
Nominal diameter, in
Screwed
1
2
1
2
Flanged
4
Valves (fully open):
Globe
14
8.2
6.9
5.7
Gate
0.30
0.24
0.16
0.11
Swing check
5.1
2.9
2.1
2.0
Angle
9.0
4.7
2.0
1.0
Elbows:
45° regular
0.39
0.32
0.30
0.29
45° long radius
90° regular
2.0
1.5
0.95
0.64
90° long radius
1.0
0.72
0.41
0.23
180° regular
2.0
1.5
0.95
0.64
180° long radius
Tees:
Line flow
0.90
0.90
0.90
0.90
Branch flow
2.4
1.8
1.4
1.1
1
2
4
8
20
13
0.80
2.0
4.5
8.5
0.35
2.0
2.4
6.0
0.16
2.0
2.0
5.8
0.07
2.0
2.0
5.5
0.03
2.0
2.0
0.21
0.50
0.40
0.41
0.40
0.20
0.39
0.30
0.35
0.30
0.19
0.30
0.19
0.30
0.21
0.16
0.26
0.15
0.25
0.15
0.14
0.21
0.10
0.20
0.10
0.24
1.0
0.19
0.80
0.14
0.64
0.10
0.58
0.07
0.41
the values for 90° elbows in Table 6.5, indicating smoother walls and/or better
design. One may conclude that (1) Table 6.5 data are probably conservative and
(2) loss factors are highly dependent on actual design and manufacturing factors,
with Table 6.5 serving only as a rough guide.
The valve losses in Table 6.5 are for the fully open condition. Losses can be
much higher for a partially open valve. Figure 6.18b gives average losses for three
0.34
Legend
Plastic elbow
Metal elbow no. 1
Metal elbow no. 2
0.32
+10%
0.30
K factor
0.28
Fig. 6.18a Recent measured loss
­coefficients for 90° elbows. These
values are less than those reported
in Table 6.5. (Coffield, R. D., P. T.
Mc Keown, and R.B. Hammond.
“Irrecoverable Pressure Loss
Coefficients for Two Elbows in
Series with Various Orientation
Angles and Separation Distances.”
Report WAPD-T-3117. Bettis Atomic
Power Laboratory, 1997.)
Curve-fit correlation
K = 1.49 Re– 0.145
0.26
–10%
0.24
0.22
0.20
0.18
0.16
0.05
0.1
0.2
0.3
0.5
1.0
Reynolds number (millions)
2.0
3.0 4.0
392
Chapter 6 Viscous Flow in Ducts
20.00
18.00
Gate
16.00
Disk
Globe
14.00
12.00
K 10.00
8.00
6.00
4.00
2.00
Fig. 6.18b Average loss coefficients
for partially open valves (see
sketches in Fig. 6.17).
0.00
0.25
0.30
0.40
0.50
0.60
0.70
0.75
0.80
0.90
1.00
Fractional opening h
D
1000.00
100.00
K 10.00
1.00
20
0.10
80
30
40
50
60
70
90
Valve opening angle, degrees
(b)
(a)
Fig. 6.19 Performance of butterfly
valves: (a) typical geometry (Courtesy of Nadezda Murmakova/Alamy
Stock Photo); (b) loss coefficients
for three different manufacturers.
valves as a function of “percentage open,” as defined by the opening-distance
ratio h/D (see Fig. 6.17 for the geometries). Again we should warn of a possible
uncertainty of ±50 percent. Of all minor losses, valves, because of their complex
geometry, are most sensitive to manufacturers’ design details. For more accuracy,
the particular design and manufacturer should be consulted [35].
The butterfly valve of Fig. 6.19a is a stem-mounted disk that, when closed,
seats against an O-ring or compliant seal near the pipe surface. A single 90° turn
opens the valve completely, hence the design is ideal for controllable quick-opening and quick-closing situations such as occur in fire protection and the electric
power industry. However, considerable dynamic torque is needed to close these
valves, and losses are high when the valves are nearly closed.
Figure 6.19b shows butterfly-valve loss coefficients as a function of the opening angle θ for turbulent flow conditions (θ = 0 is closed). The losses are huge
when the opening is small, and K drops off nearly exponentially with the opening
6.9 Minor or Local Losses in Pipe Systems 393
Secondary
flow pattern:
=
18
0°
1.0
θ
0.8
0.6
K
R
θ=
0.4
90°
5°
θ=4
d = constant
0.2
Fig. 6.20 Resistance coefficients for
smooth-walled 45°, 90°, and 180°
bends, at Red = 200,000, after Ito [49].
Source: After H. Ito, “Pressure Losses in
Smooth Pipe Bends,” Journal of Basic
­Engineering, March 1960, pp. 131–143.
0
0
5
R
d
10
15
angle. There is a factor of 2 spread among the various manufacturers. Note that
K in Fig. 6.19b is, as usual, based on the average pipe velocity V = Q/A, not on
the increased velocity of the flow as it passes through the narrow valve passage.
A bend or curve in a pipe, as in Fig. 6.20, always induces a loss larger than
the simple straight-pipe Moody friction loss, due to flow separation on the curved
walls and a swirling secondary flow arising from the centripetal acceleration. The
smooth-wall loss coefficients K in Fig. 6.20, from the data of Ito [49], are for
total loss, including Moody friction effects. The separation and secondary flow
losses decrease with R/d, while the Moody losses increase because the bend
length increases. The curves in Fig. 6.20 thus show a minimum where the two
effects cross. Ito [49] gives a curve-fit formula for the 90° bend in turbulent flow:
R 0.84
R −1.96
90° bend: K ≈ 0.388α ( ) Re−0.17
where
α
=
0.95
+
4.42
≥ 1 (6.77)
D
(d)
d
The formula accounts for Reynolds number, which equals 200,000 in Fig. 6.20.
Comprehensive reviews of curved-pipe flow, for both laminar and turbulent flows,
are given by Berger et al. [53] and for 90° bends by Spedding et al. [54].
As shown in Fig. 6.21, entrance losses are highly dependent on entrance geometry, but exit losses are not. Sharp edges or protrusions in the entrance cause large
zones of flow separation and large losses. A little rounding goes a long way, and
a well-rounded entrance (r = 0.2d) has a nearly negligible loss K = 0.05. At a
submerged exit, on the other hand, the flow simply passes out of the pipe into
the large downstream reservoir and loses all its velocity head due to viscous dissipation. Therefore K = 1.0 for all submerged exits, no matter how well rounded.
If the entrance is from a finite reservoir, it is termed a sudden contraction (SC)
between two sizes of pipe. If the exit is to finite-sized pipe, it is termed a sudden
expansion (SE). The losses for both are graphed in Fig. 6.22. For the sudden
394
Chapter 6 Viscous Flow in Ducts
1.0
t
=0
d
t
K
V
0.02
l
0.5
(a)
0
0.1
0.2
l
d
0.3
0.4
0.6
Sharp-edged
L
V
r
0.4
d
θ
K
Fig. 6.21 Entrance and exit loss
­coefficients: (a) reentrant inlets;
(b) rounded and beveled inlets. Exit
losses are K ≈ 1.0 for all shapes of
exit (reentrant, sharp, beveled, or
rounded).
θ=
10°
50°
0.2
30°
0
(b)
0
0.10
0.15
0.20
r, L
d d
Source: ASHRAE Handbook-2012
­Fundamentals, ASHRAE, 2012.
1.0
Sudden expansion
0.8
d
V
D
hm
K=
V 2/(2g)
0.6
Eq. (6.78)
Eq. (6.79)
0.4
Sudden contraction:
Vena contracta
V
0.2
d
D
Fig. 6.22 Sudden expansion and
contraction losses. Note that the
loss is based on velocity head in the
small pipe.
0
0.2
0.4
d
D
0.6
0.8
1.0
r
d
L
d
6.9 Minor or Local Losses in Pipe Systems 395
expansion, the shear stress in the corner separated flow, or deadwater region, is
negligible, so that a control volume analysis between the expansion section and
the end of the separation zone gives a theoretical loss:
KSE = (1 −
hm
d2 2
= 2
2)
D
V /(2g)
(6.78)
Note that K is based on the velocity head in the small pipe. Equation (6.78) is in
excellent agreement with experiment.
For the sudden contraction, however, flow separation in the downstream pipe
causes the main stream to contract through a minimum diameter dmin, called the
vena contracta, as sketched in Fig. 6.22. Because the theory of the vena contracta
is not well developed, the loss coefficient in the figure for sudden contraction is
experimental. It fits the empirical formula
d2
KSC ≈ 0.42 (1 − 2 )
D
(6.79)
up to the value d/D = 0.76, above which it merges into the sudden-expansion
prediction, Eq. (6.78).
Gradual Expansion—The Diffuser
As flow enters a gradual expansion or diffuser, such as the conical geometry of
Fig. 6.23, the velocity drops and the pressure rises. An efficient diffuser reduces
V1
2θ
d1
d2
V2
1.0
d1/d2 = 0.1
0.2
0.3
0.4
0.5
0.6
0.9
0.8
0.7
0.6
K 0.5
0.4
0.3
Data:
d1/d2 = 0.33
0.2
Fig. 6.23 Flow losses in a gradual
conical expansion region, as
calculated from Gibson’s suggestion
[15, 50], Eq. (6.80), for a smooth wall.
= 0.50 [16]
0.1
0.0
0
10
20
30
Total cone angle 2θ, degrees
40
50
396
Chapter 6 Viscous Flow in Ducts
the pumping power required. Head loss can be large, due to flow separation
on the walls, if the cone angle is too great. A thinner entrance boundary layer,
as in Fig. 6.6, causes a slightly smaller loss than a fully developed inlet flow.
The flow loss is a combination of nonideal pressure recovery plus wall friction.
Some ­correlating curves are shown in Fig. 6.23. The loss coefficient K is based
on the velocity head in the inlet (small) pipe and depends upon cone angle 2θ
and the diffuser diameter ratio d1/d2. There is scatter in the reported data [15,
16]. The curves in Fig. 6.23 are based on a correlation by A. H. Gibson [50],
cited in Ref. 15:
Kdiffuser =
hm
V21∕(2 g)
≈ 2.61 sin θ (1 −
d2 2
L
+ favg
2)
davg
D
for 2θ ≤ 45°(6.80)
For large angles, 2θ > 45°, drop the coefficient (2.61 sin θ), which leaves us with
a loss equivalent to the sudden expansion of Eq. (6.78). As seen, the formula is
in reasonable agreement with the data from Ref. 16. The minimum loss lies in
the region 5° < 2θ < 15°, which is the best geometry for an efficient diffuser.
For angles less than 5°, the diffuser is too long and has too much friction. Angles
greater than 15° cause flow separation, resulting in poor pressure recovery. Professor Gordon H
­ olloway provided the writer a recent example, where an improved
diffuser design reduced the power requirement of a wind tunnel by 40 percent
(100 hp decrease!). We shall look again at diffusers in Sec. 6.11, using the data
of Ref. 14.
For a gradual contraction, the loss is very small, as seen from the following
experimental values [15]:
Contraction cone angle 2θ, deg
K for gradual contraction
30
45
60
0.02
0.04
0.07
References 15, 16, 43, and 46 contain additional data on minor losses.
EXAMPLE 6.16
Water, ρ = 1.94 slugs/ft3 and ν = 0.000011 ft2/s, is pumped between two reservoirs at
0.2 ft3/s through 400 ft of 2-in-diameter pipe and several minor losses, as shown in
Fig. E6.16. The roughness ratio is ε/d = 0.001. Compute the pump horsepower required.
Solution
Write the steady flow energy equation between sections 1 and 2, the two reservoir
­surfaces:
p2 V22
p1 V21
+
+ z1 = ( +
+ z2) + hf +
ρg 2g
ρg 2g
∑ hm − hp
where hp is the head increase across the pump. But since p1 = p2 and V1 = V2 ≈ 0,
solve for the pump head:
V2 fL
hp = z2 − z1 + hf + ∑ hm = 120 ft − 20 ft + ( + ∑ K)
(1)
2g d
6.9 Minor or Local Losses in Pipe Systems 397
Screwed
regular
90° elbow
Sharp
exit
2
z2 = 120 ft
1
z1 = 20 ft
Sharp
entrance
Open globe
valve
Half-open
gate valve
12-in
bend radius
Pump
400 ft of pipe, d =
2
ft
12
E6.16
Now with the flow rate known, calculate
V=
Q
0.2 ft3/s
= 1 2 2 = 9.17 ft/s
A 4π( 12 ft)
Now list and sum the minor loss coefficients:
Loss
K
Sharp entrance (Fig. 6.21)
Open globe valve (2 in, Table 6.5)
12-in bend (Fig. 6.20)
Regular 90° elbow (Table 6.5)
Half-closed gate valve (from Fig. 6.18b)
Sharp exit (Fig. 6.21)
0.5
6.9
0.25
0.95
3.8
1.0
Σ K = 13.4
Calculate the Reynolds number and pipe friction factor:
Red =
9.17( 122 )
Vd
=
= 139,000
ν
0.000011
For ε/d = 0.001, from the Moody chart read f = 0.0216. Substitute into Eq. (1):
hp = 100 ft +
(9.17 ft /s) 2 0.0216(400)
+ 13.4 ]
2
2(32.2 ft /s2 ) [
12
= 100 ft + 85 ft = 185 ft pump head
The pump must provide a power to the water of
P = ρgQhp = [1.94(32.2) lbf/ft3 ] (0.2 ft3/s) (185 ft) ≈ 2300 ft · lbf/s
398
Chapter 6 Viscous Flow in Ducts
The conversion factor is 1 hp = 550 ft · lbf/s. Therefore
2300
P=
= 4.2 hp
Ans.
550
Allowing for an efficiency of 70 to 80 percent, a pump is needed with an input of
about 6 hp.
Laminar Flow Minor Losses
The data in Table 6.5 are for turbulent flow in fittings. If the flow is laminar,
a different form of loss occurs, which is proportional to V, not V 2. By analogy with Eqs. (6.12) for Poiseuille flow, the laminar minor loss takes the
form
Klam =
Δploss d
μV
Laminar minor losses are just beginning to be studied, due to increased interest
in micro- and nano-flows in tubes. They can be substantial, comparable to the
Poiseuille loss.
Laminar Minor Loss Coefficients Klam in Tube Fittings for 1 ≤ Red ≤ 10 can
be found in Ref. 47.
6.10 Multiple-Pipe Systems5
If you can solve the equations for one-pipe systems, you can solve them all; but
when systems contain two or more pipes, certain basic rules make the calculations
very smooth. Any resemblance between these rules and the rules for handling
electric circuits is not coincidental.
Figure 6.24 shows three examples of multiple-pipe systems.
Pipes in Series
The first is a set of three (or more) pipes in series. Rule 1 is that the flow rate
is the same in all pipes:
or
Q1 = Q2 = Q3 = const
(6.81)
V1d21 = V2d22 = V3d23
(6.82)
Rule 2 is that the total head loss through the system equals the sum of the head
loss in each pipe:
ΔhA→B = Δh1 + Δh2 + Δh3
5
This section may be omitted without loss of continuity.
(6.83)
6.10 Multiple-Pipe Systems 399
3
2
1
A
B
(a)
1
2
B
A
3
(b)
z2
HGL
z1
Fig. 6.24 Examples of multiplepipe systems: (a) pipes in series;
(b) pipes in parallel; (c) the threereservoir junction problem.
HGL
zJ +
pJ
ρg
z3
HGL
2
3
1
(c)
In terms of the friction and minor losses in each pipe, we could rewrite this as
ΔhA→B =
V21 f1L1
V22 f2L2
∑
+
K
+
+ ∑ K2)
1)
2g ( d1
2g ( d2
+
V23 f3L3
+ ∑ K3)
2g ( d3
(6.84)
and so on for any number of pipes in the series. Since V2 and V3 are proportional
to V1 from Eq. (6.82), Eq. (6.84) is of the form
ΔhA→B =
V21
(α0 + α1 f1 + α2 f2 + α3 f3 )
2g
(6.85)
where the αi are dimensionless constants. If the flow rate is given, we can
e­ valuate the right-hand side and hence the total head loss. If the head loss
is given, a little iteration is needed, since f1, f2, and f3 all depend on V1
through the Reynolds number. Begin by calculating f1, f2, and f3, assuming
fully rough flow, and the solution for V 1 will converge with one or two
iterations.
400
Chapter 6 Viscous Flow in Ducts
EXAMPLE 6.17
Given is a three-pipe series system, as in Fig. 6.24a. The total pressure drop is pA − pB =
150,000 Pa, and the elevation drop is zA − zB = 5 m. The pipe data are
Pipe
L, m
d, cm
ε, mm
ε/d
8
6
4
0.24
0.12
0.20
0.003
0.002
0.005
1
100
2
150
3 80
The fluid is water, ρ = 1000 kg/m3 and ν = 1.02 × 10−6 m2/s. Calculate the flow rate
Q in m3/h through the system.
Solution
The total head loss across the system is
p A − pB
150,000
ΔhA→B =
+ 5 m = 20.3 m
+ zA − zB =
ρg
1000(9.81)
From the continuity relation (6.85) the velocities are
V2 =
and
Re2 =
d21
16
V = V1
2 1
9
d2
V3 =
d21
d23
V2d2
4
Re1 = Re1
V1d1
3
V1 = 4V1
Re3 = 2Re1
Neglecting minor losses and substituting into Eq. (6.84), we obtain
ΔhA→B =
or
V21
16 2
2
1250f
+
2500
1
( 9 ) f2 + 2000(4) f3 ]
2g [
20.3 m =
V21
(1250f1 + 7900f2 + 32,000f3 )
2g
(1)
This is the form that was hinted at in Eq. (6.85). It seems to be dominated by the third
pipe loss 32,000f3. Begin by estimating f1, f2, and f3 from the Moody-chart fully rough
regime:
f1 = 0.0262 f2 = 0.0234 f3 = 0.0304
Substitute in Eq. (1) to find V21 ≈ 2g(20.3)/(33 + 185 + 973). The first estimate thus is
V1 = 0.58 m/s, from which
Re1 ≈ 45,400 Re2 = 60,500 Re3 = 90,800
Hence, from the Moody chart,
f1 = 0.0288 f2 = 0.0260 f3 = 0.0314
Substitution into Eq. (1) gives the better estimate
V1 = 0.565 m/s
or
Q = 14πd21V1 = 2.84 × 10−3 m3/s
Q = 10.2 m3/h
3
A second iteration gives Q = 10.22 m /h, a negligible change.
Ans.
6.10 Multiple-Pipe Systems 401
Pipes in Parallel
The second multiple-pipe system is the parallel flow case shown in Fig. 6.24b.
Here the pressure drop is the same in each pipe, and the total flow is the sum of
the individual flows:
(6.86a)
ΔhA→B = Δh1 = Δh2 = Δh3
Q = Q1 + Q2 + Q3(6.86b)
If the total head loss is known, it is straightforward to solve for Qi in each pipe
and sum them, as will be seen in Example 6.18. The reverse problem, of determining ΣQi when hf is known, requires iteration. Each pipe is related to hf by the
Moody relation hf = f(L/d)(V2/2g) = fQ2/C, where C = π2gd5/8L. Thus each pipe
has nearly quadratic nonlinear parallel resistance, and head loss is related to total
flow rate by
hf =
Q2
(Σ √Ci/fi )
2
where Ci =
π2gdi5
8Li
(6.87)
Since the fi vary with Reynolds number and roughness ratio, one begins Eq. (6.87)
by guessing values of fi (fully rough values are recommended) and calculating a first
estimate of hf. Then each pipe yields a flow-rate estimate Qi ≈ (Cihf /fi)1/2 and hence
a new Reynolds number and a better estimate of fi. Then repeat Eq. (6.87) to convergence.
It should be noted that both of these parallel-pipe cases—finding either ΣQ or
hf —are easily solved by Excel if reasonable guesses are given.
EXAMPLE 6.18
Assume that the same three pipes in Example 6.17 are now in parallel with the same
total head loss of 20.3 m. Compute the total flow rate Q, neglecting minor losses.
Solution
From Eq. (6.86a) we can solve for each V separately:
20.3 m =
V23
V21
V22
1250f1 =
2500f2 =
2000f3
2g
2g
2g
(1)
Guess fully rough flow in pipe 1: f1 = 0.0262, V1 = 3.49 m/s; hence Re1 = V1d1/ν =
273,000. From the Moody chart read f1 = 0.0267; recompute V1 = 3.46 m/s, Q1 = 62.5
m3/h.
Next guess for pipe 2: f2 ≈ 0.0234, V2 ≈ 2.61 m/s; then Re2 = 153,000, and hence
f2 = 0.0246, V2 = 2.55 m/s, Q2 = 25.9 m3/h.
Finally guess for pipe 3: f3 ≈ 0.0304, V3 ≈ 2.56 m/s; then Re3 = 100,000, and hence
f3 = 0.0313, V3 = 2.52 m/s, Q3 = 11.4 m3/h.
This is satisfactory convergence. The total flow rate is
Q = Q1 + Q2 + Q3 = 62.5 + 25.9 + 11.4 = 99.8 m3/h
Ans.
These three pipes carry 10 times more flow in parallel than they do in series.
This example may be solved by Excel iteration using the Colebrook-formula procedure outlined in Ex. 6.9. Each pipe is a separate iteration of friction factor, Reynolds
402
Chapter 6 Viscous Flow in Ducts
number, and flow rate. The pipes are rough, so only one iteration is needed. Here are
the Excel results:
A
B
Re1
(ε/d )1
1
2
313053
271100
0.003
0.003
Re2
(ε/d)2
1
2
166021
149739
0.002
0.002
Re3
(ε/d)3
1
123745
2 98891
0.005
0.005
C
D
Ex. 6.18 − Pipe 1
V1 − m/s
Q1 − m3/h
3.991
3.457
72.2
62.5
Ex. 6.18 − Pipe 2
V2 − m/s
Q2 − m3/h
2.822
2.546
28.7
25.9
Ex. 6.18 − Pipe 3
V3 − m/s
Q3 − m3/h
3.155
2.522
14.3
11.4
E
F
f1
f1-guess
0.0267
0.0267
0.0200
0.0267
f2
f2-guess
0.0246
0.0246
0.0200
0.0246
f3
f3-guess
0.0313
0.0313
0.0200
0.0313
Thus, as in the hand calculations, the total flow rate = 62.5 + 25.9 + 11.4 = 99.8 m3/h.
Ans.
Three-Reservoir Junction
Consider the third example of a three-reservoir pipe junction, as in Fig. 6.24c. If
all flows are considered positive toward the junction, then
Q1 + Q2 + Q3 = 0
(6.88)
which obviously implies that one or two of the flows must be away from the
junction. The pressure must change through each pipe so as to give the same
static pressure pJ at the junction. In other words, let the HGL at the junction have
the elevation
hJ = zJ +
pJ
ρg
where pJ is in gage pressure for simplicity. Then the head loss through each,
assuming p1 = p2 = p3 = 0 (gage) at each reservoir surface, must be such that
Δh1 =
V21 f1L1
= z1 − hJ
2g d1
Δh2 =
V22 f2L2
= z2 − hJ (6.89)
2g d2
Δh3 =
V23 f3L3
= z3 − hJ
2g d3
6.10 Multiple-Pipe Systems 403
We guess the position hJ and solve Eqs. (6.89) for V1, V2, and V3 and hence Q1,
Q2, and Q3, iterating until the flow rates balance at the junction according to
Eq. (6.88). If we guess hJ too high, the sum Q1 + Q2 + Q3 will be negative and
the remedy is to reduce hJ, and vice versa.
EXAMPLE 6.19
Take the same three pipes as in Example 6.17, and assume that they connect three
reservoirs at these surface elevations
z1 = 20 m z2 = 100 m z3 = 40 m
Find the resulting flow rates in each pipe, neglecting minor losses.
Solution
As a first guess, take hJ equal to the middle reservoir height, z3 = hJ = 40 m. This saves
one calculation (Q3 = 0) and enables us to get the lay of the land:
Reservoir
hJ, m
zi − hJ, m
fi
Vi, m/s
Qi, m3/h
1
40
−20
0.0267
−3.43
−62.1
2
40
60
0.0241 4.42
45.0
3
40
0 0
0
ΣQ = −17.1
Li/di
1250
2500
2000
Since the sum of the flow rates toward the junction is negative, we guessed hJ too high.
Reduce hJ to 30 m and repeat:
Reservoir
hJ, m
zi − hJ, m
fi
Vi, m/s
Qi, m3/h
1
30
−10
0.0269
−2.42
−43.7
2
30
70
0.0241
4.78
48.6
3
30
10
0.0317
1.76
8.0
ΣQ = 12.9
This is positive ΣQ, and so we can linearly interpolate to get an accurate guess: hJ ≈ 34.3 m.
Make one final list:
Reservoir
hJ, m
zi − hJ, m
fi
Vi, m/s
Qi, m3/h
1
34.3
−14.3
0.0268
−2.90
−52.4
2
34.3
65.7
0.0241
4.63
47.1
3 34.3
5.7
0.0321
1.32
6.0
ΣQ = 0.7
This is close enough; hence we calculate that the flow rate is 52.4 m3/h toward reservoir
3, balanced by 47.1 m3/h away from reservoir 1 and 6.0 m3/h away from reservoir 3.
One further iteration with this problem would give hJ = 34.53 m, resulting in
Q1 = −52.8, Q2 = 47.0, and Q3 = 5.8 m3/h, so that ΣQ = 0 to three-place accuracy.
Pedagogically speaking, we would then be exhausted.
404
Chapter 6 Viscous Flow in Ducts
1
2
A
B
C
5
3
Loop I
Loop II
4
E
7
6
F
10
D
8
Loop III
Loop IV
I
12
9
11
G
H
Fig. 6.25 Schematic of a piping
­network.
Pipe Networks
The ultimate case of a multipipe system is the piping network illustrated in
Fig. 6.25. This might represent a water supply system for an apartment or
­subdivision or even a city. This network is quite complex algebraically but
follows the same basic rules:
1. The net flow into any junction must be zero.
2. The net pressure change around any closed loop must be zero. In other
words, the HGL at each junction must have one and only one elevation.
3. All pressure changes must satisfy the Moody and minor-loss friction
correlations.
By supplying these rules to each junction and independent loop in the network, one
obtains a set of simultaneous equations for the flow rates in each pipe leg and the
HGL (or pressure) at each junction. Solution may then be obtained by numerical
iteration, as first developed in a hand calculation technique by Prof. Hardy Cross in
1936 [17]. Computer solution of pipe network problems is now quite common and
is covered in at least one specialized text [18]. Network analysis is quite useful for
real water distribution systems if well calibrated with the actual system head loss data.
6.11 Experimental Duct Flows: Diffuser Performance6
The Moody chart is such a great correlation for tubes of any cross section with
any roughness or flow rate that we may be deluded into thinking that the world
of internal flow prediction is at our feet. Not so. The theory is reliable only for
6
This section may be omitted without loss of continuity.
6.11 Experimental Duct Flows: Diffuser Performance 405
ducts of constant cross section. As soon as the section varies, we must rely
­
principally on experiment to determine the flow properties. As mentioned many
times before, experimentation is a vital part of fluid mechanics.
Literally thousands of papers in the literature report experimental data for specific internal and external viscous flows. We have already seen several examples:
1.
2.
3.
4.
5.
6.
Vortex shedding from a cylinder (Fig. 5.1).
Drag of a sphere and a cylinder (Fig. 5.2).
Hydraulic model of a dam spillway (Fig. 5.8).
Rough-wall pipe flows (Fig. 6.12).
Secondary flow in ducts (Fig. 6.16).
Minor duct loss coefficients (Sec. 6.9).
Chapter 7 will treat a great many more external flow experiments, especially in
Sec. 7.6. Here we shall show data for one type of internal flow, the diffuser.
Diffuser Performance
A diffuser, shown in Fig. 6.26a and b, is an expansion or area increase intended
to reduce velocity in order to recover the pressure head of the flow. Rouse and
Ince [6] relate that it may have been invented by customers of the early Roman
100
b
1
70
2
2θ
W2
L
(a)
Bistable
steady stall
c
b
40
2 θ , degrees
W1
c
Jet flow
a
20
Transitory
stall
b
Maximum
unsteadiness
10
7
Cp max
4
L
No
stall
2
2θ
D
Throat
(b)
De
Exit
1
1
2
4
a
7 10
L
W1
20
40
(c)
Fig. 6.26 Diffuser geometry and typical flow regimes: (a) geometry of a flat-walled
­diffuser; (b) geometry of a conical diffuser; (c) flat diffuser stability map. (Runstadler,
Peter W., and Francis X. Dolan, “Diffuser Data Book.” Creare Inc., 1975.)
100
406
Chapter 6 Viscous Flow in Ducts
(about 100 a.d.) water supply system, where water flowed continuously and was
billed according to pipe size. The ingenious customers discovered that they could
increase the flow rate at no extra cost by flaring the outlet section of the pipe.
Engineers have always designed diffusers to increase pressure and reduce kinetic
energy of ducted flows, but until about 1950, diffuser design was a combination of
art, luck, and vast amounts of empiricism. Small changes in design parameters caused
large changes in performance. The Bernoulli equation seemed highly suspect as a
useful tool.
Neglecting losses and gravity effects, the incompressible Bernoulli equation
­predicts that
p + 12ρV2 = p0 = const
(6.90)
where p0 is the stagnation pressure the fluid would achieve if the fluid were
slowed to rest (V = 0) without losses.
The basic output of a diffuser is the pressure-recovery coefficient Cp, defined as
pe − pt
Cp =
(6.91)
p0t − pt
where subscripts e and t mean the exit and the throat (or inlet), respectively.
Higher Cp means better performance.
Consider the flat-walled diffuser in Fig. 6.26a, where section 1 is the inlet
and section 2 the exit. Application of Bernoulli’s equation (6.90) to this diffuser ­predicts that
p01 = p1 + 12ρV21 = p2 + 12ρV22 = p02
or
Cp,frictionless = 1 − (
V2 2
V1 )
(6.92)
Meanwhile, steady one-dimensional continuity would require that
Q = V1A1 = V2A2
(6.93)
Combining (6.92) and (6.93), we can write the performance in terms of the area
ratio AR = A2/A1, which is a basic parameter in diffuser design:
Cp,frictionless = 1 − (AR) −2
(6.94)
A typical design would have AR = 5:1, for which Eq. (6.94) predicts Cp = 0.96,
or nearly full recovery. But, in fact, measured values of Cp for this area ratio [14]
are only as high as 0.86 and can be as low as 0.24.
The basic reason for the discrepancy is flow separation, as sketched in
Fig. 6.27b. The increasing pressure in the diffuser is an unfavorable gradient
(Sec. 7.5), which causes the viscous boundary layers to break away from the
walls and greatly reduces the performance. Computational fluid dynamics (CFD)
can now predict this behavior.
As an added complication to boundary layer separation, the flow patterns in
a diffuser are highly variable and were considered mysterious and erratic until
1955, when Kline revealed the structure of these patterns with flow visualization
techniques in a simple water channel.
6.11 Experimental Duct Flows: Diffuser Performance 407
Thin
boundary
layers
Low
velocity,
high
pressure
(a)
Backflow
Thick
boundary
layers
High
velocity,
low
pressure
Fig. 6.27 Diffuser performance:
(a) ideal pattern with good performance; (b) actual measured pattern
with boundary layer separation and
resultant poor performance.
Separation
point
“Stalled”
flow
(b)
A complete stability map of diffuser flow patterns was published in 1962
by Fox and Kline [21], as shown in Fig. 6.26c. There are four basic regions.
Below line aa there is steady viscous flow, no separation, and moderately
good performance. Note that even a very short diffuser will separate, or stall,
if its half-angle is greater than 10°.
Between lines aa and bb is a transitory stall pattern with strongly unsteady
flow. Best performance (highest Cp) occurs in this region. The third pattern,
between bb and cc, is steady bistable stall from one wall only. The stall pattern
may flip-flop from one wall to the other, and performance is poor.
The fourth pattern, above line cc, is jet flow, where the wall separation is so
gross and pervasive that the mainstream ignores the walls and simply passes on
through at nearly constant area. Performance is extremely poor in this region.
Dimensional analysis of a flat-walled or conical diffuser shows that Cp should
depend on the following parameters:
1. Any two of the following geometric parameters:
a. Area ratio AR = A2/A1 or (De/D)2
b. Divergence angle 2θ
c. Slenderness L/W1 or L/D
408
Chapter 6 Viscous Flow in Ducts
2. Inlet Reynolds number Ret = V1W1/ν or Ret = V1D/ν
3. Inlet Mach number Mat = V1/a1
4. Inlet boundary layer blockage factor Bt = ABL/A1, where ABL is the wall
area blocked, or displaced, by the retarded boundary layer flow in the inlet
(typically Bt varies from 0.03 to 0.12)
A flat-walled diffuser would require an additional shape parameter to describe
its cross section:
5. Aspect ratio AS = b/W1
Even with this formidable list, we have omitted five possible important effects:
inlet turbulence, inlet swirl, inlet profile vorticity, superimposed pulsations,
and downstream obstruction, all of which occur in practical machinery
­applications.
The three most important parameters are AR, θ, and B. Typical performance
maps for diffusers are shown in Fig. 6.28. For this case of 8 to 9 percent ­blockage,
AS
Mat
Bt
ReD
= 1.0
= 0.2
= 0.08
= 279,000
h
Flat
5
Transitory
stall
boundary
4.5
Cp
0.7
0
4
3.5
0.69
0.68
AR
3
0.66
0.64
20°
2
1.75
Fig. 6.28a Typical performance
maps for flat-wall and conical
­diffusers at similar operating
­conditions: flat wall. (Runstadler,
Peter W., and Francis X. Dolan,
“Diffuser Data Book.” Creare Inc.,
1975.)
0.62
18°
0.60
16°
14°
12°
10°
8°
4
5
2θ = 4°
6°
6
7
8 9
L
W1
(a)
10
12
14
16 18 20
6.12 Fluid Meters 409
Mt = 0.2
Bt = 0.09
Red = 120,000
Conical
25
2 θ = 18° 16° 14° 12° 10°
8°
16
6°
12
10
5°
4°
8
AR
0.70
6
3°
0.68
5
4
0.44
0.66
0.46
3
0.54
0.48
0.50 0.52
2.5
Cp
0.62
0.60
0.58
0.56
0.64
2°
2
Fig. 6.28b Typical performance
maps for flat-wall and conical
­diffusers at similar operating
­conditions: conical wall.
(Runstadler, Peter W., and Francis
X. Dolan, “Diffuser Data Book.”
Creare Inc., 1975.)
1.75
1.5
2
4
6
8
10 12
16
Diffuser length–throat diameter ratio L
d
(b)
20
25 30
both the flat-walled and conical types give about the same maximum ­performance,
Cp = 0.70, but at different divergence angles (9° flat versus 4.5° conical). Both
types fall far short of the Bernoulli estimates of Cp = 0.93 (flat) and 0.99 ­(conical),
primarily because of the blockage effect.
The experimental design of a diffuser is an excellent example of a successful
attempt to minimize the undesirable effects of adverse pressure gradient and flow
separation.
6.12 Fluid Meters
Almost all practical fluid engineering problems are associated with the need for
an accurate flow measurement. There is a need to measure local properties (velocity, pressure, temperature, density, viscosity, turbulent intensity), integrated properties (mass flow and volume flow), and global properties (visualization of the
entire flow field). We shall concentrate in this section on velocity and volume
flow measurements.
410
Chapter 6 Viscous Flow in Ducts
We have discussed pressure measurement in Sec. 2.10. Measurement of other
thermodynamic properties, such as density, temperature, and viscosity, is beyond
the scope of this text and is treated in specialized books such as Refs. 22 and 23.
Global visualization techniques were discussed in Sec. 1.11 for low-speed flows,
and the special optical techniques used in high-speed flows are treated in Ref.
34 of Chap. 1. Flow measurement schemes suitable for open-channel and other
free-surface flows are treated in Chap. 10.
Local Velocity Measurements
Velocity averaged over a small region, or point, can be measured by several
different physical principles, listed in order of increasing complexity and
­
­sophistication:
1. Trajectory of floats or neutrally buoyant particles.
2. Rotating mechanical devices:
a. Cup anemometer.
b. Savonius rotor.
c. Propeller meter.
d. Turbine meter.
3. Pitot-static tube (Fig. 6.30).
4. Electromagnetic current meter.
5. Hot wires and hot films.
6. Laser-doppler anemometer (LDA).
7. Particle image velocimetry (PIV).
Some of these meters are sketched in Fig. 6.29.
Floats or Buoyant Particles. A simple but effective estimate of flow velocity can
be found from visible particles entrained in the flow. Examples include flakes on
the surface of a channel flow, small neutrally buoyant spheres mixed with a ­liquid,
or hydrogen bubbles. Sometimes gas flows can be estimated from the motion of
entrained dust particles. One must establish whether the particle motion truly
simulates the fluid motion. Floats are commonly used to track the movement of
ocean waters and can be designed to move at the surface, along the bottom, or
at any given depth [24]. Many official tidal current charts [25] were obtained by
releasing and timing a floating spar attached to a length of string. One can release
whole groups of spars to determine a flow pattern.
Rotating Sensors. The rotating devices of Fig. 6.29a to d can be used in either
gases or liquids, and their rotation rate is approximately proportional to the flow
velocity. The cup anemometer (Fig. 6.29a) and Savonius rotor (Fig. 6.29b) always
rotate the same way, regardless of flow direction. They are popular in atmospheric
and oceanographic applications and can be fitted with a direction vane to align
themselves with the flow. The ducted-propeller (Fig. 6.29c) and free-propeller
(Fig. 6.29d) meters must be aligned with the flow parallel to their axis of rotation.
They can sense reverse flow because they will then rotate in the opposite ­direction.
6.12 Fluid Meters 411
( b)
(a)
(c)
Plated film
Fine
wire
( d)
(e)
(f )
Display
Fig. 6.29 Eight common velocity
meters: (a) three-cup anemometer;
(b) Savonius rotor; (c) turbine
mounted in a duct; (d) free-propeller
meter; (e) hot-wire anemometer;
(f) hot-film anemometer; (g) pitotstatic tube; (h) laser-doppler
­anemometer.
θ
Laser
( g)
Focusing
optics
Flow
( h)
Receiving Photo
optics detector
All these rotating sensors can be attached to counters or sensed by electromagnetic or slip-ring devices for either a continuous or a digital reading of flow
velocity. All have the disadvantage of being relatively large and thus not representing a “point.”
Pitot-Static Tube. A slender tube aligned with the flow (Figs. 6.29g and 6.30)
can measure local velocity by means of a pressure difference. It has sidewall holes
to measure the static pressure ps in the moving stream and a hole in the front to
measure the stagnation pressure p0, where the stream is decelerated to zero velocity. Instead of measuring p0 or ps separately, it is customary to measure their
difference with, say, a transducer, as in Fig. 6.30.
If ReD > 1000, where D is the probe diameter, the flow around the probe is
nearly frictionless and Bernoulli’s relation, Eq. (3.54), applies with good accuracy. For incompressible flow
ps + 12ρV2 + ρgzs ≈ p0 + 12ρ(0) 2 + ρgz0
412
Chapter 6 Viscous Flow in Ducts
Static ≈ Free-stream
pressure
pressure
8D
V
θ
ps
Stagnation
pressure
4 to 8
holes
Error
+10%
Fig. 6.30 Pitot-static tube for
­combined measurement of static
and stagnation pressure in a
­moving stream.
–10%
p0
Static pressure
0
pS
Stagnation pressure
0°
10°
Yaw angle θ
Differential
pressure
transducer
20°
Assuming that the elevation pressure difference ρg(zs − z0) is negligible, this
reduces to
V ≈ [2
(p0 − ps ) 1/2
] ρ
(6.95)
This is the Pitot formula, named after the French engineer Henri de Pitot, who
designed the device in 1732.
The primary disadvantage of the pitot tube is that it must be aligned with the
flow direction, which may be unknown. For yaw angles greater than 5°, there are
substantial errors in both the p0 and ps measurements, as shown in Fig. 6.30. The
pitot-static tube is useful in liquids and gases; for gases a compressibility correction
is necessary if the stream Mach number is high (Chap. 9). Because of the slow
response of the fluid-filled tubes leading to the pressure sensors, it is not useful
for unsteady flow measurements. It does resemble a point and can be made small
enough to measure, for example, blood flow in arteries and veins. It is not suitable
for low-velocity measurement in gases because of the small pressure differences
developed. For example, if V = 1 ft/s in standard air, from Eq. (6.95) we compute
p0 − p equal to only 0.001 lbf/ft2 (0.048 Pa). This is beyond the resolution of most
pressure gages.
Electromagnetic Meter. If a magnetic field is applied across a conducting fluid,
the fluid motion will induce a voltage across two electrodes placed in or near the
flow. The electrodes can be streamlined or built into the wall, and they cause
little or no flow resistance. The output is very strong for highly conducting fluids
such as liquid metals. Seawater also gives good output, and electromagnetic current meters are in common use in oceanography. Even low-conductivity fresh
water can be measured by amplifying the output and insulating the electrodes.
Commercial instruments are available for most liquid flows but are relatively
costly. Electromagnetic flowmeters are treated in Ref. 26.
6.12 Fluid Meters 413
Hot-Wire Anemometer. A very fine wire (d = 0.01 mm or less) heated between
two small probes, as in Fig. 6.29e, is ideally suited to measure rapidly fluctuating
flows such as the turbulent boundary layer. The idea dates back to work by L. V.
King in 1914 on heat loss from long, thin cylinders. If electric power is supplied
to heat the cylinder, the loss varies with flow velocity across the cylinder according to King’s law
q = I2R ≈ a + b(ρV) n
(6.96)
1
2
1
3
where n ≈ at very low Reynolds numbers and equals at high Reynolds numbers. The hot wire normally operates in the high-Reynolds-number range but
should be calibrated in each situation to find the best-fit a, b, and n. The wire
can be operated either at constant current I, so that resistance R is a measure of
V, or at constant resistance R (constant temperature), with I a measure of velocity. In either case, the output is a nonlinear function of V, and the equipment
should contain a linearizer to produce convenient velocity data. Many varieties
of commercial hot-wire equipment are available, as are do-it-yourself designs
[27]. Excellent detailed discussions of the hot wire are given in Ref. 28.
Because of its frailty, the hot wire is not suited to liquid flows, whose high
density and entrained sediment will knock the wire right off. A more stable yet
quite sensitive alternative for liquid flow measurement is the hot-film anemometer (Fig. 6.29f). A thin metallic film, usually platinum, is plated onto a relatively
thick support, which can be a wedge, a cone, or a cylinder. The operation is
similar to the hot wire. The cone gives best response but is liable to error when
the flow is yawed to its axis.
Hot wires can easily be arranged in groups to measure two- and three-dimensional velocity components.
Laser-Doppler Anemometer. In the LDA a laser beam provides highly focused,
coherent monochromatic light that is passed through the flow. When this light is
scattered from a moving particle in the flow, a stationary observer can detect a
change, or doppler shift, in the frequency of the scattered light. The shift Δf is
proportional to the velocity of the particle. There is essentially zero disturbance
of the flow by the laser.
Figure 6.29h shows the popular dual-beam mode of the LDA. A focusing
device splits the laser into two beams, which cross the flow at an angle θ. Their
intersection, which is the measuring volume or resolution of the measurement,
resembles an ellipsoid about 0.5 mm wide and 0.1 mm in diameter. Particles
passing through this measuring volume scatter the beams; they then pass through
receiving optics to a photodetector, which converts the light to an electric signal.
A signal processor then converts electric frequency to a voltage that can be either
displayed or stored. If λ is the wavelength of the laser light, the measured velocity is given by
V=
λ Δf
2 sin (θ/2)
(6.97)
Multiple components of velocity can be detected by using more than one photodetector and other operating modes. Either liquids or gases can be measured as
414
Chapter 6 Viscous Flow in Ducts
long as scattering particles are present. In liquids, normal impurities serve as
scatterers, but gases may have to be seeded. The particles may be as small as the
wavelength of the light. Although the measuring volume is not as small as with
a hot wire, the LDA is capable of measuring turbulent fluctuations.
The advantages of the LDA are as follows:
1.
2.
3.
4.
5.
No disturbance of the flow.
High spatial resolution of the flow field.
Velocity data that are independent of the fluid thermodynamic properties.
An output voltage that is linear with velocity.
No need for calibration.
The disadvantages are that both the apparatus and the fluid must be transparent
to light and that the cost is high (a basic system shown in Fig. 6.29h begins at
about $50,000).
Once installed, an LDA can map the entire flow field in minutest detail. To
truly appreciate the power of the LDA, one should examine, for instance, the
amazingly detailed three-dimensional flow profiles measured by Eckardt [29] in
a high-speed centrifugal compressor impeller. Extensive discussions of laser velocimetry are given in Refs. 38 and 39.
Particle Image Velocimetry. This popular new idea, called PIV for short, measures
not just a single point but instead maps the entire field of flow. An illustration
was shown in Fig. 1.19b. The flow is seeded with neutrally buoyant particles. A
planar laser light sheet across the flow is pulsed twice and photographed twice.
If Δr is the particle displacement vector over a short time Δt, an estimate of its
velocity is V ≈ Δr/Δt. A dedicated computer applies this formula to a whole
cloud of particles and thus maps the flow field. One can also use the data to
calculate velocity gradient and vorticity fields. Since the particles all look alike,
other cameras may be needed to identify them. Three-dimensional velocities can
be measured by two cameras in a stereoscopic arrangement. The PIV method is
not limited to stop-action. New high-speed cameras (up to 10,000 frames per
second) can record movies of unsteady flow fields. For further details, see the
monograph by M. Raffel [51].
EXAMPLE 6.20
The pitot-static tube of Fig. 6.30 uses mercury as a manometer fluid. When it is placed
in a water flow, the manometer height reading is h = 8.4 in. Neglecting yaw and other
errors, what is the flow velocity V in ft/s?
Solution
From the two-fluid manometer relation (2.23b), with zA = z2, the pressure difference is
related to h by
p0 − ps = (γM − γw )h
6.12 Fluid Meters 415
Taking the specific weights of mercury and water from Table 2.1, we have
p0 − ps = (846 − 62.4 lbf/ft3 )
8.4
ft = 549 lbf/ft2
12
The density of water is 62.4/32.2 = 1.94 slugs/ft3. Introducing these values into the
pitot-static formula (6.95), we obtain
2(549 lbf/ft2 ) 1/2
V=[
= 23.8 ft/s
1.94 slugs/ft3 ]
Ans.
Since this is a low-speed flow, no compressibility correction is needed.
Volume Flow Measurements
It is often desirable to measure the integrated mass, or volume flow, passing
through a duct. Accurate measurement of flow is vital in billing customers for a
given amount of liquid or gas passing through a duct. The different devices available to make these measurements are discussed in great detail in the ASME text
on fluid meters [30]. These devices split into two classes: mechanical instruments
and head loss instruments.
The mechanical instruments measure actual mass or volume of fluid by trapping it and counting it. The various types of measurement are
1. Mass measurement
a. Weighing tanks
b. Tilting traps
2. Volume measurement
a. Volume tanks
b. Reciprocating pistons
c. Rotating slotted rings
d. Nutating disc
e. Sliding vanes
f. Gear or lobed impellers
g. Reciprocating bellows
h. Sealed-drum compartments
The last three of these are suitable for gas flow measurement.
The head loss devices obstruct the flow and cause a pressure drop, which is
a measure of flux:
1. Bernoulli-type devices
a. Thin-plate orifice
b. Flow nozzle
c. Venturi tube
2. Friction loss devices
a. Capillary tube
b. Porous plug
416
Chapter 6 Viscous Flow in Ducts
E
D
C
Fig. 6.31 Cutaway sketch of
a nutating disc fluid meter.
A: metered-volume chamber;
B: nutating disc; C: rotating spindle;
D: drive magnet; E: magnetic
­counter sensor.
B
Source: Courtesy of Badger Meter, Inc.,
Milwaukee, Wisconsin.
A
The friction loss meters cause a large nonrecoverable head loss and obstruct the
flow too much to be generally useful.
Six other widely used meters operate on different physical principles:
1.
2.
3.
4.
5.
6.
Turbine meter
Vortex meter
Ultrasonic flowmeter
Rotameter
Coriolis mass flowmeter
Laminar flow element
Nutating Disc Meter. For measuring liquid volumes, as opposed to volume rates,
the most common devices are the nutating disc and the turbine meter. Figure 6.31
shows a cutaway sketch of a nutating disc meter, widely used in both water and
gasoline delivery systems. The mechanism is clever and perhaps beyond the
writer’s capability to explain. The metering chamber is a slice of a sphere and
contains a rotating disc set at an angle to the incoming flow. The fluid causes
the disc to nutate (spin eccentrically), and one revolution corresponds to a certain
fluid volume passing through. Total volume is obtained by counting the number
of revolutions.
Turbine Meter. The turbine meter, sometimes called a propeller meter, is a freely
rotating propeller that can be installed in a pipeline. A typical design is shown
in Fig. 6.32a. There are flow straighteners upstream of the rotor, and the rotation
is measured by electric or magnetic pickup of pulses caused by passage of a point
on the rotor. The rotor rotation is approximately proportional to the volume flow
in the pipe.
Like the nutating disc, a major advantage of the turbine meter is that each
pulse corresponds to a finite incremental volume of fluid, and the pulses are
digital and can be summed easily. Liquid flow turbine meters have as few as two
blades and produce a constant number of pulses per unit fluid volume over a 5:1
6.12 Fluid Meters 417
Magnetic
pulse
pickup
Turbine
rotor
Rotor supports
Percent registration
Fig. 6.32 The turbine meter widely
used in the oil and gas industry:
(a) basic design; (b) the linearity
curve is the measure of variation in
the signal output across the 10% to
100% nominal flow range of the
meter. (Daniel Measurement and
Control, Houston, TX.)
Indicated linearity
(a)
Meter range at the indicated linearity
101.
± 0.15%
100.
99.
0
10
20
30
40
50
60
Flow rate %
70
80
90
100
flow rate range with ±0.25 percent accuracy. Gas meters need many blades to
produce sufficient torque and are accurate to ±1 percent.
Since turbine meters are very individualistic, flow calibration is an absolute
necessity. A typical liquid meter calibration curve is shown in Fig. 6.32b.
Researchers attempting to establish universal calibration curves have met with
little practical success as a result of manufacturing variabilities.
Turbine meters can also be used in unconfined flow situations, such as winds
or ocean currents. They can be compact, even microsize with two or three component directions. Figure 6.33 illustrates a handheld wind velocity meter that uses
a seven-bladed turbine with a calibrated digital output. The accuracy of this
device is quoted at ±2 percent.
418
Chapter 6 Viscous Flow in Ducts
Fig. 6.33 A Commercial handheld
wind velocity turbine meter.
(Courtesy of Nielsen-Kellerman.)
Vortex Flowmeters. Recall from Fig. 5.1 that a bluff body placed in a uniform
crossflow sheds alternating vortices at a nearly uniform Strouhal number
St = fL/U, where U is the approach velocity and L is a characteristic body width.
Since L and St are constant, this means that the shedding frequency is proportional to velocity:
f = (const)(U)
(6.98)
The vortex meter introduces a shedding element across a pipe flow and picks up
the shedding frequency downstream with a pressure, ultrasonic, or heat transfer
type of sensor. A typical design is shown in Fig. 6.34.
The advantages of a vortex meter are as follows:
1.
2.
3.
4.
5.
Absence of moving parts.
Accuracy to ±1 percent over a wide flow rate range (up to 100:1).
Ability to handle very hot or very cold fluids.
Requirement of only a short pipe length.
Calibration insensitive to fluid density or viscosity.
For further details see Ref. 40.
Ultrasonic Flowmeters. The sound-wave analog of the laser velocimeter of
Fig. 6.29h is the ultrasonic flowmeter. Two examples are shown in Fig. 6.35. The
pulse-type flowmeter is shown in Fig. 6.35a. Upstream piezoelectric transducer
A is excited with a short sonic pulse that propagates across the flow to downstream transducer B. The arrival at B triggers another pulse to be created at A,
resulting in a regular pulse frequency fA. The same process is duplicated in the
reverse direction from B to A, creating frequency fB. The difference fA − fB is
proportional to the flow rate. Figure 6.35b shows a doppler-type arrangement,
where sound waves from transmitter T are scattered by particles or contaminants
in the flow to receiver R. Comparison of the two signals reveals a doppler
6.12 Fluid Meters 419
Sensor
Force on sensor
Pivoting axis
Shedder bar
Electronic measurement module
Fluid flow
Vortex shedder force
Piezoelectric
frequency sensor
Fig. 6.34 A vortex flowmeter.
A
B
(a)
R
(b)
T
(c)
Fig. 6.35 Ultrasonic flowmeters: (a) pulse type; (b) doppler-shift type (a-b: Vass, Gabor.
“Ultrasonic Flowmeter Basics.” Sensors 14, no. 10 (1997): 73–78.); (c) a clamp-on
noninvasive installation (Watcharapol Amprasert/123RF)
420
Chapter 6 Viscous Flow in Ducts
f­requency shift that is proportional to the flow rate. Ultrasonic meters are nonintrusive and can be directly attached to pipe flows in the field (Fig. 6.35c). Their
quoted uncertainty of ±1 to 2 percent can rise to ±5 percent or more due to
irregularities in velocity profile, fluid temperature, or Reynolds number. For further details see Ref. 41.
Rotameter. The variable-area transparent rotameter of Fig. 6.36 has a float that,
under the action of flow, rises in the vertical tapered tube and takes a certain
equilibrium position for any given flow rate. A student exercise for the forces on
the float would yield the approximate relation
Q = Cd Aa (
1/2
2Wnet
Afloat ρfluid )
(6.99)
where Wnet is the float’s net weight in the fluid, Aa = Atube − Afloat is the annular
area between the float and the tube, and Cd is a dimensionless discharge coefficient
of order unity, for the annular constricted flow. For slightly tapered tubes, Aa varies
nearly linearly with the float position, and the tube may be calibrated and marked
with a flow rate scale, as in Fig. 6.36. The rotameter thus provides a readily visible
measure of the flow rate. Capacity may be changed by using different-sized floats.
Obviously the tube must be vertical, and the device does not give accurate readings
for fluids containing high concentrations of bubbles or particles.
Coriolis Mass Flowmeter. Most commercial meters measure volume flow, with
Fig. 6.36 A commercial rotameter.
The float rises in the tapered tube to
an equilibrium position, which is a
measure of the fluid flow rate.
(Courtesy of Blue White Industries.)
mass flow then computed by multiplying by the nominal fluid density. An attractive ­modern alternative is a mass flowmeter, which operates on the principle of
the Coriolis acceleration associated with noninertial coordinates [recall Fig. 3.11
and the Coriolis term 2Ω × V in Eq. (3.48)]. The output of the meter is directly
proportional to mass flow.
Figure 6.37 is a schematic of a Coriolis device, to be inserted into a piping
system. The flow enters a loop arrangement, which is electromagnetically vibrated
Flow out
Flow in
Fluid force
Fig. 6.37 Schematic of a Coriolis
mass flowmeter.
Fluid force
Vibrating tube
6.12 Fluid Meters 421
Self-sealing pressure
measurement connection
O-ring-sealed pressure connection
Platinum resistance
thermometer
Microprocessor
Piston-centering seat
Pressure-equalization chamber
Annular laminar-flow
path defined
by piston
and cylinder
Fig. 6.38 A complete flowmeter system using a laminar flow element (in this case a narrow
annulus). The flow rate is linearly proportional to the pressure drop.
at a high natural frequency (amplitude < 1 mm and frequency > 100 Hz). The
Coriolis effect induces a downward force on the loop entrance and an upward
force on the loop exit, as shown. The loop twists, and the twist angle can be
measured and is proportional to the mass flow through the tube. Accuracy is
typically less than 1 percent of full scale.
Laminar Flow Element. In many, perhaps most, commercial flowmeters, the flow
through the meter is turbulent and the variation of flow rate with pressure drop
is nonlinear. In laminar duct flow, however, Q is linearly proportional to Δp, as
in Eq. (6.12): Q = [πR4/(8µL)] Δp. Thus a laminar flow sensing element is attractive, since its calibration will be linear. To ensure laminar flow for what otherwise
would be a turbulent condition, all or part of the fluid is directed into small passages, each of which has a low (laminar) Reynolds number. A honeycomb is a
popular design.
Figure 6.38 uses axial flow through a narrow annulus to create laminar flow.
The theory again predicts Q∝Δp, as in Eq. (6.70). However, the flow is very
sensitive to passage size; for example, halving the annulus clearance increases Δp
more than eight times. Careful calibration is thus necessary. In Fig. 6.38 the
laminar flow concept has been synthesized into a complete mass flow system,
with temperature control, differential pressure measurement, and a microprocessor
all self-contained. The accuracy of this device is rated at ±0.2 percent.
422
Chapter 6 Viscous Flow in Ducts
Horizontal
EGL
Moody
loss
p1 – p2
HGL
Nonrecoverable
head loss
Vena contracta D2
D
V1
d =β D
V2 ≈ V1
( )
D
D2
2
Dividing
streamline
Fig. 6.39 Velocity and pressure
change through a generalized
­Bernoulli obstruction meter.
Deadwater
region
Bernoulli Obstruction Theory. Consider the generalized flow obstruction shown
in Fig. 6.39. The flow in the basic duct of diameter D is forced through an
obstruction of diameter d; the β ratio of the device is a key parameter:
d
(6.100)
D
After leaving the obstruction, the flow may neck down even more through a vena
contracta of diameter D2 < d, as shown. Apply the Bernoulli and continuity equations for incompressible steady frictionless flow to estimate the pressure change:
β=
Continuity:
Bernoulli:
π
π
Q = D2V1 = D22V2
4
4
1
2
p0 = p1 + 2ρV1 = p2 + 12ρV22
Eliminating V1, we solve these for V2 or Q in terms of the pressure change p1 − p2:
2(p1 − p2 ) 1/2
Q
= V2 ≈ [
A2
ρ(1 − D42/D4 ) ]
(6.101)
But this is surely inaccurate because we have neglected friction in a duct flow,
where we know friction will be very important. Nor do we want to get into the
6.12 Fluid Meters 423
business of measuring vena contracta ratios D2/d for use in (6.101). Therefore we
assume that D2/D ≈ β and then calibrate the device to fit the relation
Q = AtVt = Cd At[
2(p1 − p2 )/ρ
4
1−β
]
1/2
(6.102)
where subscript t denotes the throat of the obstruction. The dimensionless discharge coefficient Cd accounts for the discrepancies in the approximate analysis.
By dimensional analysis for a given design we expect
Cd = f (β, ReD )
where ReD =
V1D
ν
(6.103)
The geometric factor involving β in (6.102) is called the velocity-of-approach
factor:
E = (1 − β4 ) −1/2
(6.104)
One can also group Cd and E in Eq. (6.102) to form the dimensionless flow coefficient α:
α = CdE =
Cd
(1 − β4 ) 1/2
(6.105)
Thus Eq. (6.102) can be written in the equivalent form
Q = αAt[
2(p1 − p2 ) 1/2
] ρ
(6.106)
Obviously the flow coefficient is correlated in the same manner:
α = f (β, ReD )
(6.107)
Occasionally one uses the throat Reynolds number instead of the approach Reynolds number:
Red =
Vt d ReD
=
ν
β
(6.108)
Since the design parameters are assumed known, the correlation of α from Eq. (6.107)
or of Cd from Eq. (6.103) is the desired solution to the fluid metering problem.
The mass flow is related to Q by

m = ρQ
(6.109)
and is thus correlated by exactly the same formulas.
Figure 6.40 shows the three basic devices recommended for use by the
­International Organization for Standardization (ISO) [31]: the orifice, nozzle, and
venturi tube.
Thin-Plate Orifice. The thin-plate orifice, Fig. 6.40b, can be made with β in the
range of 0.2 to 0.8, except that the hole diameter d should not be less than 12.5
mm. To measure p1 and p2, three types of tappings are commonly used:
424
Chapter 6 Viscous Flow in Ducts
Ellipse
2d
3
d
Bevel angle:
45° to 60°
0.6 d
d
Flow
D
Flow
d
Edge thickness:
0.005 D to 0.02 D
t2 < 13 mm
Plate thickness:
up to 0.05 D
t1 < 0.15 D
(b)
(a)
ISA 1932
nozzle shape
Fig. 6.40 Standard shapes for the
three primary Bernoulli obstructiontype meters: (a) long-radius nozzle;
(b) thin-plate orifice; (c) venturi
nozzle. (Based on data from the
­International Organization for
­Standardization.)
Conical
diffuser
Throat tap
D
2
Flow
θ < 15°
0.7d
d
2
(c)
1. Corner taps where the plate meets the pipe wall.
2. D: 12 D taps: pipe-wall taps at D upstream and 12 D downstream.
3. Flange taps: 1 in (25 mm) upstream and 1 in (25 mm) downstream of the
plate, regardless of the size D.
Types 1 and 2 approximate geometric similarity, but since the flange taps 3 do
not, they must be correlated separately for every single size of pipe in which a
flange-tap plate is used [30, 31].
Figure 6.41 shows the discharge coefficient of an orifice with D: 12 D or type
2 taps in the Reynolds number range ReD = 104 to 107 of normal use. Although
detailed charts such as Fig. 6.41 are available for designers [30], the ASME
recommends use of the curve-fit formulas developed by the ISO [31]. The basic
form of the curve fit is [42]
Cd = f (β) + 91.71β2.5Re−0.75
+
D
where
0.09β4
1 − β4
F1 − 0.0337β3F2
f (β) = 0.5959 + 0.0312β2.1 − 0.184β8
(6.110)
6.12 Fluid Meters 425
0.66
0.65
D
0.7
β = 0.8 = d
D
0.64
0.63
Cd
1
D
2
p2
p1
Flow
d
D
0.6
0.62
0.5
0.61
0.60
0.4
0.3
0.2
0.59
Fig. 6.41 Discharge coefficient for a
thin-plate orifice with D: 12 D taps,
plotted from Eqs. (6.110) and
(6.111b).
0.58
10 4
105
10 6
107
ReD
The correlation factors F1 and F2 vary with tap position:
Corner taps:
F1 = 0 F2 = 0
(6.111a)
D: 12D taps:
F1 = 0.4333 F2 = 0.47
(6.111b)
Flange taps:
1
F2 =
D (in)
1
F1 = • D (in)
0.4333
D > 2.3 in
2.0 ≤ D ≤ 2.3 in
(6.111c)
Note that the flange taps (6.111c), not being geometrically similar, use raw diameter in inches in the formula. The constants will change if other diameter units
are used. We cautioned against such dimensional formulas in Example 1.4 and
give Eq. (6.111c) only because flange taps are widely used in the United States.
Flow Nozzle. The flow nozzle comes in two types, a long-radius type shown in
Fig. 6.40a and a short-radius type (not shown) called the ISA 1932 nozzle [30,
31]. The flow nozzle, with its smooth, rounded entrance convergence, practically
eliminates the vena contracta and gives discharge coefficients near unity. The
426
Chapter 6 Viscous Flow in Ducts
nonrecoverable loss is still large because there is no diffuser provided for gradual
expansion.
The ISO recommended correlation for long-radius-nozzle discharge coefficient is
Cd ≈ 0.9965 − 0.00653β1/2 (
106 1/2
106 1/2
=
0.9965
−
0.00653
( Red ) (6.112)
ReD )
The second form is independent of the β ratio and is plotted in Fig. 6.42. A
similar ISO correlation is recommended for the short-radius ISA 1932 flow nozzle:
Cd ≈ 0.9900 − 0.2262β4.1
+ (0.000215 − 0.001125β + 0.00249β4.7 ) (
106 1.15
ReD )
(6.113)
Flow nozzles may have β values between 0.2 and 0.8.
Venturi Meter. The third and final type of obstruction meter is the venturi,
named in honor of Giovanni Venturi (1746–1822), an Italian physicist who first
tested conical expansions and contractions. The original, or classical, venturi was
invented by a U.S. engineer, Clemens Herschel, in 1898. It consisted of a 21°
conical contraction, a straight throat of diameter d and length d, then a 7° to 15°
conical expansion. The discharge coefficient is near unity, and the nonrecoverable
loss is very small. Herschel venturis are seldom used now.
The modern venturi nozzle, Fig. 6.40c, consists of an ISA 1932 nozzle entrance
and a conical expansion of half-angle no greater than 15°. It is intended to be
operated in a narrow Reynolds number range of 1.5 × 105 to 2 × 106. Its discharge
coefficient, shown in Fig. 6.43, is given by the ISO correlation formula
Cd ≈ 0.9858 − 0.196β4.5
(6.114)
It is independent of ReD within the given range. The Herschel venturi discharge
varies with ReD but not with β, as shown in Fig. 6.42. Both have very low net losses.
1.00
0.99
0.98
Cd
Classical
Herschel venturi (ReD)
0.97
0.96
0.95
0.94
Fig. 6.42 Discharge coefficient for
long-radius nozzle and classical
Herschel-type venturi.
0.93
104
All values
of β
Long-radius
nozzle (Red )
105
106
Red , ReD
107
108
6.12 Fluid Meters 427
1.00
0.98
Cd
0.96
0.94
Fig. 6.43 Discharge coefficient for a
venturi nozzle.
0.92
0.3
International
standards:
0.316 < β < 0.775
1.5 × 10 5 < ReD < 2.0 × 106
0.4
0.5
β
0.6
0.7
0.8
The choice of meter depends on the loss and the cost and can be illustrated
by the following table:
Type of meter
Net head loss
Cost
Orifice
Nozzle
Venturi
Large
Medium
Small
Small
Medium
Large
As so often happens, the product of inefficiency and initial cost is approximately
constant.
The average nonrecoverable head losses for the three types of meters, expressed
as a fraction of the throat velocity head Vt2/(2g), are shown in Fig. 6.44. The orifice has the greatest loss and the venturi the least, as discussed. The orifice and
nozzle simulate partially closed valves as in Fig. 6.18b, while the venturi is a very
minor loss. When the loss is given as a fraction of the measured pressure drop,
the orifice and nozzle have nearly equal losses, as Example 6.21 will illustrate.
The other types of instruments discussed earlier in this section can also serve
as flowmeters if properly constructed. For example, a hot wire mounted in a tube
can be calibrated to read volume flow rather than point velocity. Such hot-wire
meters are commercially available, as are other meters modified to use velocity
instruments. For further details see Ref. 30.
Compressible Gas Flow Correction Factor. The orifice/nozzle/venturi formulas in
this section assume incompressible flow. If the fluid is a gas, and the pressure
ratio (p2/p1) is not near unity, a compressibility correction is needed. Equation
(6.101) is rewritten in terms of mass flow and the upstream density ρ1:
2ρ1 (p1 − p2 )

m = Cd Y At √
1 − β4
where β =
d
D
(6.115)
428
Chapter 6 Viscous Flow in Ducts
3.0
2.5
Thin-plate
orifice
hm
1.5
Km =
Vt2/(2g)
2.0
1.0
Flow
nozzle
Venturi:
0.5
Fig. 6.44 Nonrecoverable head loss
in Bernoulli obstruction meters.
(Adapted from Ref. 30.)
0
15° cone angle
7° cone angle
0.2
0.3
0.4
0.5
0.6
0.7
0.8
β
The dimensionless expansion factor Y is a function of pressure ratio, β, and the
type of meter. Some values are plotted in Fig. 6.45. The orifice, with its strong
jet contraction, has a different factor from the venturi or the flow nozzle, which
are designed to eliminate contraction.
1
Sharp-edged orifices:
β = 0.2 0.5 0.7 0.8
Expansion factor, Y
0.9
0.8
β = 0.2 0.5 0.6 0.7 0.8
Nozzles and venturis:
0.7
Fig. 6.45 Compressible flow expansion factor Y for flowmeters.
0.6
0.6
0.7
0.8
p2 /p1
0.9
1
6.12 Fluid Meters 429
EXAMPLE 6.21
We want to meter the volume flow of water (ρ = 1000 kg/m3, ν = 1.02 × 10−6 m2/s)
moving through a 200-mm-diameter pipe at an average velocity of 2.0 m/s. If the differential pressure gage selected reads accurately at p1 − p2 = 50,000 Pa, what size
meter should be selected for installing (a) an orifice with D: 12 D taps, (b) a long-radius
flow nozzle, or (c) a venturi nozzle? What would be the nonrecoverable head loss for
each design?
Solution
Here the unknown is the β ratio of the meter. Since the discharge coefficient is a complicated function of β, iteration will be necessary. We are given D = 0.2 m and V1 = 2.0 m/s.
The pipe-approach Reynolds number is thus
ReD =
V1D
(2.0) (0.2)
=
= 392,000
v
1.02 × 10−6
For all three cases [(a) to (c)] the generalized formula (6.105) holds:
Vt =
V1
β2
= α[
2(p1 − p2 ) 1/2
Cd
α=
]
ρ
(1 − β4 ) 1/2
(1)
where the given data are V1 = 2.0 m/s, ρ = 1000 kg/m3, and Δp = 50,000 Pa. Inserting
these known values into Eq. (1) gives a relation between β and α:
2(50,000) 1/2
2.0
= α[
or
2
1000 ]
β
β2 =
0.2
α
(2)
The unknowns are β (or α) and Cd. Parts (a) to (c) depend on the particular chart or
formula needed for Cd = fcn(ReD, β). We can make an initial guess β ≈ 0.5 and iterate to
convergence.
Part (a)
For the orifice with D: 12 D taps, use Eq. (6.110) or Fig. 6.41. The iterative sequence is
β1 ≈ 0.5, Cd1 ≈ 0.604, α1 ≈ 0.624, β2 ≈ 0.566, Cd2 ≈ 0.606, α2 ≈ 0.640, β3 = 0.559
We have converged to three figures. The proper orifice diameter is
d = βD = 112 mm
Ans. (a)
Part (b)
For the long-radius flow nozzle, use Eq. (6.112) or Fig. 6.42. The iterative sequence is
β1 ≈ 0.5, Cd1 ≈ 0.9891, α1 ≈ 1.022, β2 ≈ 0.442, Cd2 ≈ 0.9896, α2 ≈ 1.009, β3 = 0.445
We have converged to three figures. The proper nozzle diameter is
d = βD = 89 mm
Ans. (b)
430
Chapter 6 Viscous Flow in Ducts
Part (c)
For the venturi nozzle, use Eq. (6.114) or Fig. 6.43. The iterative sequence is
β1 ≈ 0.5, Cd1 ≈ 0.977, α1 ≈ 1.009, β2 ≈ 0.445, Cd2 ≈ 0.9807, α2 ≈ 1.0004, β3 = 0.447
We have converged to three figures. The proper venturi diameter is
d = βD = 89 mm
Ans. (c)
Comments: These meters are of similar size, but their head losses are not the same.
From Fig. 6.44 for the three different shapes we may read the three K factors and compute
hm,orifice ≈ 3.5 m hm,nozzle ≈ 3.6 m hm,venturi ≈ 0.8 m
The venturi loss is only about 22 percent of the orifice and nozzle losses.
Solution by Excel Iteration for the Flow Nozzle
Parts (a, b, c) were solved by hand, but Excel is ideal for these calculations. You may review
this procedure from the instructions in Example 6.5. We need five columns: Cd, calculated
from Eq. (6.112), throat velocity Vt calculated from Δp, α as calculated from Eq. (6.105),
and β calculated from the velocity ratio (V/Vt). The fifth column is an initial guess for β,
which is replaced in its next row by the newly computed β. Any initial β < 1 will do. Here
we chose β = 0.5 as in part (b) for the flow nozzle. Remember to use cell names, not symbols: in row 1, Cd = A1, Vt = B1, α = C1, and β = D1. The process converges rapidly, in
only two or three iterations:
Cd from Eq.(6.114)
Vt = α(2Δp/ρ)
α=
Cd/(1 − β 4)0.5
β=
(V/Vt)0.5
β-guess
A
B
C
D
E
1
2
3
4
0.9891
0.9896
0.9895
0.9895
10.216
10.091
10.096
10.096
1.0216
1.0091
1.0096
1.0096
0.4425
0.4452
0.4451
0.4451
0.5000
0.4425
0.4452
0.4451
The final answers for the long-radius flow nozzle are:
α = 1.0096 Cd = 0.9895 β = 0.4451
Ans. (b)
EXAMPLE 6.22
A long-radius nozzle of diameter 6 cm is used to meter airflow in a 10-cm-diameter
pipe. Upstream conditions are p1 = 200 kPa and T1 = 100°C. If the pressure drop
through the nozzle is 60 kPa, estimate the flow rate in m3/s.
Solution
∙ Assumptions: The pressure drops 30 percent, so we need the compressibility factor
Y, and Eq. (6.115) is applicable to this problem.
Summary 431
∙ Approach: Find ρ1 and Cd and apply Eq. (6.115) with β = 6/10 = 0.6.
∙ Property values: Given p1 and T1, ρ1 = p1/RT1 = (200,000)/[287(100 + 273)] =
1.87 kg/m3. The downstream pressure is p2 = 200 − 60 = 140 kPa, hence p2/p1 =
0.7. At 100°C, from Table A.2, the viscosity of air is 2.17 E−5 kg/m-s.
∙ Solution steps: Initially apply Eq. (6.115) by guessing, from Fig. 6.42, that Cd ≈
0.98. From Fig. 6.45, for a nozzle with p2/p1 = 0.7 and β = 0.6, read Y ≈ 0.80. Then
2(1.87 kg/m3 ) (60,000 Pa)
2ρ1 (p1 − p2 )
π

2
m = Cd YAt √
≈
(0.98)
(0.80)
(0.06
m)
√
4
1 − (0.6)
1 − β4
≈ 1.13
kg
s
Now estimate Red, putting it in the convenient mass flow form:

4(1.13 kg/s)
ρVd 4 m
Red =
=
=
≈ 1.11 E6
μ
πμd π(2.17 E−5 kg/m · s) (0.06 m)
Returning to Fig. 6.42, we could read a slightly better Cd ≈ 0.99. Thus our final
­estimate is

m ≈ 1.14 kg/s
Ans.
∙Comments: Figure 6.45 is not just a “chart” for engineers to use casually. It is based
on the compressible flow theory of Chap. 9. There, we may reassign this example as
a theory.
Summary
This chapter has been concerned with internal pipe and duct flows, which are
probably the most common problems encountered in engineering fluid mechanics.
Such flows are very sensitive to the Reynolds number and change from laminar
to transitional to turbulent flow as the Reynolds number increases.
The various Reynolds number regimes were outlined, and a semiempirical
approach to turbulent flow modeling was presented. The chapter then made a
detailed analysis of flow through a straight circular pipe, leading to the famous
Moody chart (Fig. 6.13) for the friction factor. Possible uses of the Moody
chart were discussed for flow rate and sizing problems, as well as the application of the Moody chart to noncircular ducts using an equivalent duct “diameter.” The addition of minor losses due to valves, elbows, fittings, and other
devices was presented in the form of loss coefficients to be incorporated along
with Moody-type friction losses. Multiple-pipe systems were discussed briefly
and were seen to be quite complex algebraically and appropriate for computer
solution.
Diffusers are added to ducts to increase pressure recovery at the exit of a
system. Their behavior was presented as experimental data, since the theory of
real diffusers is still not well developed. The chapter ended with a discussion of
flowmeters, especially the pitot-static tube and the Bernoulli obstruction type of
meter. Flowmeters also require careful experimental calibration.
432
Chapter 6 Viscous Flow in Ducts
Problems
Most of the problems herein are fairly straightforward. More
difficult or open-ended assignments are labeled with an asterisk. Problems labeled with a computer icon
may require
the use of a computer. The standard end-of-chapter problems
P6.1 to P6.163 (categorized in the problem list here) are followed by word problems W6.1 to W6.4, fundamentals of engineering exam problems FE6.1 to FE6.15, comprehensive
problems C6.1 to C6.9, and design projects D6.1 and D6.2.
P6.5
Problem Distribution
Section
6.1
6.2
6.3
6.4
6.5
6.6
6.7
6.8
6.9
6.10
6.10
6.11
6.12
6.12
6.12
6.12
6.12
6.12
Topic
Problems
Reynolds number regimes P6.1–P6.5
Internal viscous flows P6.6–P6.7
Head loss—friction factor P6.8–P6.11
Laminar pipe flow P6.12–P6.33
Turbulence modeling P6.34–P6.40
Turbulent pipe flow P6.41–P6.62
Flow rate and sizing problems P6.63–P6.85
Noncircular ducts P6.86–P6.98
Minor or local losses P6.99–P6.110
Series and parallel pipe systems
P6.111–P6.120
Three-reservoir and pipe network systems P6.121–P6.130
Diffuser performance
P6.131–P6.134
The pitot-static tube
P6.135–P6.139
Flowmeters: the orifice plate
P6.140–P6.148
Flowmeters: the flow nozzle
P6.149–P6.153
Flowmeters: the venturi meter
P6.154–P6.159
Flowmeters: other designs
P6.160–P6.161
Flowmeters: compressibility correction
P6.162–P6.163
Reynolds number regimes
P6.1
P6.2
P6.3
P6.4
expect transition to turbulence at (a) 20°C and (b)
100°C?
In flow past a body or wall, early transition to turbulence
can be induced by placing a trip wire on the wall across
the flow, as in Fig. P6.5. If the trip wire in Fig. P6.5 is
placed where the local velocity is U, it will trigger turbulence if Ud/ν = 850, where d is the wire diameter [3, p.
388]. If the sphere diameter is 20 cm and transition is
observed at ReD = 90,000, what is the diameter of the
trip wire in mm?
An engineer claims that the flow of SAE 30W oil, at
20°C, through a 5-cm-diameter smooth pipe at
1 million N/h, is laminar. Do you agree? A million
newtons is a lot, so this sounds like an awfully high
flow rate.
The present pumping rate of crude oil through the
Alaska Pipeline, with an ID of 48 in, is 550,000 barrels
per day (1 barrel = 42 U.S. gallons). (a) Is this a turbulent flow? (b) What would be the maximum rate if the
flow were constrained to be laminar? Assume that
Alaskan oil fits Fig. A.1 of the Appendix at 60°C.
The Keystone Pipeline in the chapter opener photo has a
maximum proposed flow rate of 1.3 million barrels of
crude oil per day. Estimate the Reynolds number and
whether the flow is laminar. Assume that Keystone
crude oil fits Fig. A.1 of the Appendix at 40°C.
For flow of SAE 30 oil through a 5-cm-diameter pipe,
from Fig. A.1, for what flow rate in m3/h would we
Trip wire d
D
U
P6.5
Internal Viscous Flows
P6.6
P6.7
Air at 68°F and 1 atm flows inside a pipe at a mass flow
rate of 0.06 lb/s. What is the minimum diameter of the
pipe if the flow is to be laminar?
SAE 10W30 oil at 20°C flows from a tank into a 2-cmdiameter tube 40 cm long. The flow rate is 1.1 m3/hr. Is
the entrance length region a significant part of this tube
flow?
Head loss––friction factor
P6.8
P6.9
x, m
p, kPa
When water at 20°C is in steady turbulent flow through
an 8-cm-diameter pipe, the wall shear stress is 72 Pa.
What is the axial pressure gradient (∂p/∂x) if the pipe is
(a) horizontal and (b) vertical with the flow up?
A light liquid (ρ ≈ 950 kg/m3) flows at an average velocity of 10 m/s through a horizontal smooth tube of
­diameter 5 cm. The fluid pressure is measured at 1-m
intervals along the pipe, as follows:
0
1
2
3
4
5
6
304
273
255
240
226
213
200
Estimate (a) the total head loss, in meters; (b) the wall
shear stress in the fully developed section of the pipe;
and (c) the overall friction factor.
P6.10 Water at 20°C flows through an inclined 8-cm-diameter
pipe. At sections A and B the following data are taken:
pA = 186 kPa, VA = 3.2 m/s, zA = 24.5 m, and pB = 260
Problems 433
kPa, VB = 3.2 m/s, zB = 9.1 m. Which way is the flow
going? What is the head loss in meters?
P6.11 Water at 20°C flows upward at 4 m/s in a 6-cm-diameter
pipe. The pipe length between points 1 and 2 is 5 m, and
point 2 is 3 m higher. A mercury manometer, connected
between 1 and 2, has a reading h = 135 mm, with p1
higher. (a) What is the pressure change (p1 − p2)? (b)
What is the head loss, in meters? (c) Is the manometer
reading proportional to head loss? Explain. (d) What is
the friction factor of the flow?
Laminar pipe flow––no minor losses
In Probs. 6.12 to 6.98, neglect minor losses.
P6.12 A 5-mm-diameter capillary tube is used as a viscometer
for oils. When the flow rate is 0.071 m3/h, the measured
pressure drop per unit length is 375 kPa/m. Estimate the
viscosity of the fluid. Is the flow laminar? Can you also
estimate the density of the fluid?
P6.13 A soda straw is 20 cm long and 2 mm in diameter. It
delivers cold cola, approximated as water at 10°C, at a
rate of 3 cm3/s. (a) What is the head loss through the
straw? What is the axial pressure gradient ∂p/∂x if the
flow is (b) vertically up or (c) horizontal? Can the human lung deliver this much flow?
P6.14 Water at 20°C is to be siphoned through a tube 1 m long
and 2 mm in diameter, as in Fig. P6.14. Is there any
height H for which the flow might not be laminar? What
is the flow rate if H = 50 cm? Neglect the tube curvature.
which has an inside diameter of 25 mm and carries
three times the flow of the smaller pipe. Both small
pipes have the same length and pressure drop. If all
flows are laminar, estimate the diameter of the smaller
pipe.
P6.17 A capillary viscometer measures the time required for a
specified volume υ of liquid to flow through a smallbore glass tube, as in Fig. P6.17. This transit time is then
correlated with fluid viscosity. For the system shown,
(a) derive an approximate formula for the time required,
assuming laminar flow with no entrance and exit losses.
(b) If L = 12 cm, l = 2 cm, υ = 8 cm3, and the fluid is
water at 20°C, what capillary diameter D will result in a
transit time t of 6 seconds?
υ
L
D
Large reservoir
L = 1 m, d = 2 mm
P6.17 Water at 20°C
H
P6.14
P6.15 Professor Gordon Holloway and his students at the
University of New Brunswick went to a fast-food emporium and tried to drink chocolate shakes (ρ ≈ 1200
kg/m3, µ ≈ 6 kg/m-s) through fat straws 8 mm in diameter and 30 cm long. (a) Verify that their human
lungs, which can develop approximately 3000 Pa of
vacuum pressure, would be u­ nable to drink the milkshake through the vertical straw. (b) A student cut
15 cm from his straw and proceeded to drink happily.
What rate of milkshake flow was produced by this
strategy?
P6.16 Fluid flows steadily, at volume rate Q, through a large
pipe and then divides into two small pipes, the larger of
P6.18 SAE 50W oil at 20°C flows from one tank to another
through a tube 160 cm long and 5 cm in diameter.
Estimate the flow rate in m 3/hr if z 1 = 2 m and
z2 = 0.8 m.
(1)
(2)
P6.18 434
Chapter 6 Viscous Flow in Ducts
P6.19 An oil (SG = 0.9) issues from the pipe in Fig. P6.19 at
Q = 35 ft3/h. What is the kinematic viscosity of the oil in
ft3/s? Is the flow laminar?
D2 = 1 cm
Q
1.5 cm
10 ft
F
3 cm
P6.22
Q
L = 6 ft
D = 1 in
2
P6.19
D1 = 0.25 mm
P6.20 The oil tanks in Tinyland are only 160 cm high, and they
discharge to the Tinyland oil truck through a smooth
tube 4 mm in diameter and 55 cm long. The tube exit is
open to the atmosphere and 145 cm below the tank surface. The fluid is medium fuel oil, ρ = 850 kg/m3 and µ
= 0.11 kg/(m · s). Estimate the oil flow rate in cm3/h.
P6.21 In Tinyland, houses are less than a foot high! The
rainfall is laminar! The drainpipe in Fig. P6.21 is only
2 mm in diameter. (a) When the gutter is full, what is
the rate of draining? (b) The gutter is designed for a
sudden rainstorm of up to 5 mm per hour. For this
condition, what is the maximum roof area that can be
drained successfully? (c) What is Red?
P6.23 SAE 10 oil at 20°C flows in a vertical pipe of diameter 2.5
cm. It is found that the pressure is constant throughout
the fluid. What is the oil flow rate in m3/h? Is the flow
up or down?
P6.24 Two tanks of water at 20°C are connected by a capillary
tube 4 mm in diameter and 3.5 m long. The surface of
tank 1 is 30 cm higher than the surface of tank 2. (a)
Estimate the flow rate in m3/h. Is the flow laminar? (b)
For what tube diameter will Red be 500?
P6.25 For the configuration shown in Fig. P6.25, the fluid is
ethyl alcohol at 20°C, and the tanks are very wide.
Find the flow rate which occurs in m3/h. Is the flow
laminar?
50 cm
Water
2 mm
80 cm
20 cm
Tinyland
governor’s
mansion
P6.21
P6.22 A steady push on the piston in Fig. P6.22 causes a flow rate
Q = 0.15 cm3/s through the needle. The fluid has
ρ = 900 kg/m3 and µ = 0.002 kg/(m · s). What force F is
required to maintain the flow?
40 cm
1m
P6.25
P6.26 Two oil tanks are connected by two 9-m-long pipes, as
in Fig. P6.26. Pipe 1 is 5 cm in diameter and is 6 m
higher than pipe 2. It is found that the flow rate in pipe
2 is twice as large as the flow in pipe 1. (a) What is
the diameter of pipe 2? (b) Are both pipe flows laminar? (c) What is the flow rate in pipe 2 (m3/s)? Neglect minor losses.
Problems 435
za = 22 m
Thousands
of straws
zb = 15 m
D1 = 5 cm
SAE
30 W
oil at
20°C
50
cm
6 m/s
D2
6m
L=9m
30 cm
P6.28
P6.26
*P6.27 Let us attack Prob. P6.25 in symbolic fashion, using
Fig. P6.27. All parameters are constant except the upper
tank depth Z(t). Find an expression for the flow rate Q(t)
as a function of Z(t). Set up a differential equation, and
solve for the time t0 to drain the upper tank completely.
Assume quasi-steady laminar flow.
P6.29 SAE 30W oil at 20°C flows through a straight pipe 25 m
long, with diameter 4 cm. The average velocity is 2 m/s.
(a) Is the flow laminar? Calculate (b) the pressure drop
and (c) the power required. (d) If the pipe diameter is
doubled, for the same average velocity, by what percent
does the required power increase?
P6.30 SAE 10 oil at 20°C flows through the 4-cm-diameter
vertical pipe of Fig. P6.30. For the mercury manometer
reading h = 42 cm shown, (a) calculate the volume flow
rate in m3/h and (b) state the direction of flow.
D
SAE 10 oil
ρ, µ
Z (t)
3m
d
h
D = 4 cm
42 cm
L
H
P6.27
P6.28 For straightening and smoothing an airflow in a 50-cm-­
diameter duct, the duct is packed with a “honeycomb” of
thin straws of length 30 cm and diameter 4 mm, as in
Fig. P6.28. The inlet flow is air at 110 kPa and 20°C,
moving at an average velocity of 6 m/s. Estimate the
pressure drop across the honeycomb.
P6.30
Mercury
P6.31 A laminar flow element (LFE) (Meriam Instrument Co.)
measures low gas-flow rates with a bundle of capillary
tubes or ducts packed inside a large outer tube. Consider
oxygen at 20°C and 1 atm flowing at 84 ft3/min in a 4-indiameter pipe. (a) Is the flow turbulent when approaching the element? (b) If there are 1000 capillary tubes, L
= 4 in, select a tube diameter to keep Red below 1500
and also to keep the tube pressure drop no greater than
0.5 lbf/in2. (c) Do the tubes selected in part (b) fit nicely
within the approach pipe?
436
Chapter 6 Viscous Flow in Ducts
P6.32 SAE 30 oil at 20°C flows in the 3-cm-diameter pipe in
Fig. P6.32, which slopes at 37°. For the pressure measurements shown, determine (a) whether the flow is up
or down and (b) the flow rate in m3/h.
pB = 180 kPa
P6.37 Two infinite plates a distance h apart are parallel to the
xz plane with the upper plate moving at speed V, as in
Fig. P6.37. There is a fluid of viscosity µ and constant
pressure between the plates. Neglecting gravity and assuming incompressible turbulent flow u(y) between the
plates, use the logarithmic law and appropriate boundary
conditions to derive a formula for dimensionless wall
shear stress versus dimensionless plate velocity. Sketch a
typical shape of the profile u(y).
u
y
20 m
V
ν
pA = 500 kPa
15 m
h
x
37°
P6.37
P6.32
P6.33 Water at 20°C is pumped from a reservoir through a vertical tube 10 ft long and 1/16th in in diameter. The pump
provides a pressure rise of 11 lbf/in2 to the flow. Neglect
entrance losses. (a) Calculate the exit velocity. (b) Approximately how high will the exit water jet rise? (c)
Verify that the flow is laminar.
Turbulence modeling
P6.34 Derive the time-averaged x-momentum equation (6.21)
by direct substitution of Eqs. (6.19) into the momentum
equation (6.14). It is convenient to write the convective
acceleration as
du
∂ 2
∂
∂
=
(u ) +
(uv) +
(uw)
dt
∂x
∂y
∂z
which is valid because of the continuity relation, Eq. (6.14).
P6.35 In the overlap layer of Fig. 6.9a, turbulent shear is large.
If we neglect viscosity, we can replace Eq. (6.24) with
the approximate velocity-gradient function
du
= fcn(y, τw, ρ)
dy
Show by dimensional analysis that this leads to the logarithmic overlap relation (6.28).
P6.36 The following turbulent flow velocity data u(y), for air at
75°F and 1 atm near a smooth flat wall were taken in the
University of Rhode Island wind tunnel:
y, in
0.025
0.035
0.047
0.055
0.065
u, ft/s
51.2
54.2
56.8
57.6
59.1
Estimate (a) the wall shear stress and (b) the velocity u at
y = 0.22 in.
Fixed
P6.38 Suppose in Fig. P6.37 that h = 3 cm, the fluid in water
at 20°C, and the flow is turbulent, so that the logarithmic law is valid. If the shear stress in the fluid is 15 Pa,
what is V in m/s?
P6.39 By analogy with laminar shear, τ = µ du/dy, T. V.
­Boussinesq in 1877 postulated that turbulent shear could
also be related to the mean velocity gradient τturb = ε du/
dy, where ε is called the eddy viscosity and is much
larger than µ. If the logarithmic overlap law, Eq. (6.28),
is valid with τturb ≈ τw, show that ε ≈ κρu*y.
P6.40 Theodore von Kármán in 1930 theorized that turbulent
shear could be represented by τturb = ε du/dy, where
ε = ρκ2y2|du/dy| is called the mixing-length eddy viscosity and κ ≈ 0.41 is Kármán’s dimensionless mixinglength constant [2, 3]. Assuming that τturb ≈ τw near the
wall, show that this expression can be integrated to yield
the logarithmic overlap law, Eq. (6.28).
Turbulent pipe flow––no minor losses
P6.41 Two reservoirs, which differ in surface elevation by 40 m,
are connected by 350 m of new pipe of diameter 8 cm. If the
desired flow rate is at least 130 N/s of water at 20°C, can the
pipe material be made of (a) galvanized iron, (b) commercial steel, or (c) cast iron? Neglect minor losses.
P6.42 Fluid flows steadily, at volume rate Q, through a large
horizontal pipe and then divides into two small pipes,
the larger of which has an inside diameter of 25 mm and
carries three times the flow of the smaller pipe. Both
small pipes have the same length and pressure drop. If all
flows are turbulent, at ReD near 104, estimate the diameter of the smaller pipe.
P6.43 A reservoir supplies water through 100 m of 30-cmdiameter cast iron pipe to a turbine that extracts 80 hp
Problems 437
from the flow. The water then exhausts to the
atmosphere.
z1 = 35 m
Cast iron pipe
Water
at 20°C
P6.48 Follow up Prob. P6.46 with the following question. If the
total Keystone pipeline length, from Alberta to Texas, is
2147 miles, how much flow, in barrels per day, will
­result if the total available pumping power is 8,000 hp?
P6.49 The tank–pipe system of Fig. P6.49 is to deliver at least
11 m3/h of water at 20°C to the reservoir. What is the
maximum roughness height ε allowable for the pipe?
z2 = 5 m
P6.43
P6.44
P6.45
P6.46
P6.47
Turbine
Neglect minor losses. (a) Assuming that f ≈ 0.019, find the
flow rate (which results in a cubic polynomial). Explain
why there are two legitimate solutions. (b) For extra credit,
solve for the flow rates using the actual friction factors.
Mercury at 20°C flows through 4 m of 7-mm-diameter
glass tubing at an average velocity of 5 m/s. Estimate the
head loss in m and the pressure drop in kPa.
Oil, SG = 0.88 and ν = 4 E−5 m2/s, flows at 400 gal/min
through a 6-in asphalted cast iron pipe. The pipe is 0.5
mi long and slopes upward at 8° in the flow direction.
Compute the head loss in ft and the pressure change.
The Keystone Pipeline in the chapter opener photo has a
diameter of 36 inches and a design flow rate of 590,000
barrels per day of crude oil at 40°C. If the pipe material
is new steel, estimate the pump horsepower required per
mile of pipe.
The gutter and smooth drainpipe in Fig. P6.47 remove
rainwater from the roof of a building. The smooth drainpipe is 7 cm in diameter. (a) When the gutter is full, estimate the rate of draining. (b) The gutter is designed for a
sudden rainstorm of up to 5 inches per hour. For this condition, what is the maximum roof area that can be drained
successfully?
Water
4m
Water at 20°C
L = 5 m, d = 3 cm
2m
P6.49
P6.50 Ethanol at 20°C flows at 125 U.S. gal/min through a
horizontal cast iron pipe with L = 12 m and d = 5 cm.
Neglecting entrance effects, estimate (a) the pressure
gradient dp/dx, (b) the wall shear stress τw, and (c) the
percentage reduction in friction factor if the pipe walls
are polished to a smooth surface.
P6.51 The viscous sublayer (Fig. 6.9) is normally less than
1 percent of the pipe diameter and therefore very difficult
to probe with a finite-sized instrument. In an effort to generate a thick sublayer for probing, Pennsylvania State University in 1964 built a pipe with a flow of glycerin.
Assume a smooth 12-in-diameter pipe with V = 60 ft/s
and glycerin at 20°C. Compute the sublayer thickness in
inches and the pumping horsepower required at 75 percent
efficiency if L = 40 ft.
P6.52 The pipe flow in Fig. P6.52 is driven by pressurized air
in the tank. What gage pressure p1 is needed to provide a
20°C water flow rate Q = 60 m3/h?
30 m
Smooth pipe:
d = 5 cm
Open jet
p1
80 m
4.2 m
10 m
P6.47
P6.52
60 m
Q
438
Chapter 6 Viscous Flow in Ducts
P6.53 Water at 20°C flows by gravity through a smooth pipe from
one reservoir to a lower one. The elevation difference is
60 m. The pipe is 360 m long, with a diameter of 12 cm.
Calculate the expected flow rate in m3/h. Neglect minor
losses.
*P6.54 A swimming pool W by Y by h deep is to be emptied by
gravity through the long pipe shown in Fig. P6.54. Assuming an average pipe friction factor fav and neglecting minor
losses, derive a formula for the time to empty the tank from
an initial level ho.
P6.59 The following data were obtained for flow of 20°C water
at 20 m3/h through a badly corroded 5-cm-diameter pipe
that slopes downward at an angle of 8°: p1 = 420 kPa, z1
= 12 m, p2 = 250 kPa, z2 = 3 m. Estimate (a) the roughness ratio of the pipe and (b) the percentage change in
head loss if the pipe were smooth and the flow rate the
same.
P6.60 In the spirit of Haaland’s explicit pipe friction factor
­approximation, Eq. (6.49), Jeppson [20] proposed the
­following explicit formula:
1
√f
Water
h
Bottom =
W by Y
Pipe:
L, D, Ɛ
V
P6.54
P6.55 The reservoirs in Fig. P6.55 contain water at 20°C. If the
pipe is smooth with L = 4500 m and d = 4 cm, what will
the flow rate in m3/h be for Δz = 100 m?
≈ −2.0 log10 (
ε/d 5.74
+
)
3.7 Re0.9
d
(a) Is this identical to Haaland’s formula with just a simple rearrangement? Explain. (b) Compare Jeppson’s formula to Haaland’s for a few representative values of
(turbulent) Red and ε/d and their errors compared to the
Colebrook formula (6.48). Discuss briefly.
P6.61 What level h must be maintained in Fig. P6.61 to deliver
a flow rate of 0.015 ft3/s through the 12 -in commercial
steel pipe?
Water
at 20°C
1
h
∆z
L = 80 ft
D = 1 in
2
L, D, Ɛ
B
2
P6.55
P6.56 The Alaska Pipeline in the chapter opener photo has a
design flow rate of 4.4 E7 gallons per day of crude oil at
60°C (see Fig. A.1). (a) Assuming a galvanized-iron
wall, estimate the total pressure drop required for the
800-mile trip. (b) If there are nine equally spaced
pumps, estimate the horsepower each pump must
­deliver.
P6.57 Apply the analysis of Prob. P6.54 to the following data.
Let W = 5 m, Y = 8 m, ho = 2 m, L = 15 m, D = 5 cm,
and ε = 0. (a) By letting h = 1.5 m and 0.5 m as representative depths, estimate the average friction factor.
Then (b) estimate the time to drain the pool.
P6.58 For the system in Prob. 6.53, a pump is used at night
to drive water back to the upper reservoir. If the pump
delivers 15,000 W to the water, estimate the flow
rate.
P6.61
P6.62 Water at 20°C is to be pumped through 2000 ft of pipe
from reservoir 1 to 2 at a rate of 3 ft3/s, as shown in Fig.
P6.62. If the pipe is cast iron of diameter 6 in and the
pump is 75 percent efficient, what horsepower pump is
needed?
120 ft
2
L = 2000 ft
P6.62
1
Pump
Problems 439
Flow rate and sizing problems
3
P6.63 A tank contains 1 m of water at 20°C and has a drawncapillary outlet tube at the bottom, as in Fig. P6.63. Find
the outlet volume flux Q in m3/h at this instant.
P6.64 For the system in Fig. P6.63, solve for the flow rate in
m3/h if the fluid is SAE 10 oil at 20°C. Is the flow laminar or turbulent?
1m
1 m3
L = 80 cm
D = 4 cm
P6.63
Q
*P6.71 It is desired to solve Prob. 6.62 for the most economical pump and cast iron pipe system. If the pump
costs $125 per horsepower delivered to the fluid and
the pipe costs $7000 per inch of diameter, what are
the minimum cost and the pipe and pump size to
maintain the 3 ft3/s flow rate? Make some simplifying assumptions.
P6.72 Modify Prob. P6.57 by letting the diameter be unknown.
Find the proper pipe diameter for which the pool will
drain in about two hours flat.
P6.73 For 20°C water flow in a smooth, horizontal 10-cm pipe,
with Δp/L = 1000 Pa/m, the writer computed a flow rate
of 0.030 m3/s. (a) Verify, or disprove, the writer’s answer. (b) If verified, use the power-law friction factor
relation, Eq. (6.41), to estimate the pipe diameter that
will triple this flow rate. (c) For extra credit, use the
more exact friction factor relation, Eq. (6.38), to solve
part (b).
P6.74 Two reservoirs, which differ in surface elevation by 40
m, are connected by a new commercial steel pipe of diameter 8 cm. If the desired flow rate is 200 N/s of water
at 20°C, what is the proper length of the pipe?
P6.75 You wish to water your garden with 100 ft of 58 -in-diameter hose whose roughness is 0.011 in. What will be the
delivery, in ft3/s, if the gage pressure at the faucet is 60
lbf/in2? If there is no nozzle (just an open hose exit),
what is the maximum horizontal distance the exit jet will
carry?
P6.76 The small turbine in Fig. P6.76 extracts 400 W of power
from the water flow. Both pipes are wrought iron. Compute the flow rate Q in m3/h. Why are there two solutions? Which is better?
P6.65 In Prob. P6.63 the initial flow is turbulent. As the water
drains out of the tank, will the flow revert to laminar motion as the tank becomes nearly empty? If so, at what
tank depth? Estimate the time, in h, to drain the tank
completely.
P6.66 Ethyl alcohol at 20°C flows through a 10-cm horizontal
drawn tube 100 m long. The fully developed wall shear
stress is 14 Pa. Estimate (a) the pressure drop, (b) the
volume flow rate, and (c) the velocity u at r = 1 cm.
P6.67 A straight 10-cm commercial-steel pipe is 1 km long and
is laid on a constant slope of 5°. Water at 20°C flows
downward, due to gravity only. Estimate the flow rate in
Water
20 m
m3/h. What happens if the pipe length is 2 km?
20°C
*P6.68 The Moody chart cannot find V directly, since V appears
in both ordinate and abscissa. (a) Arrange the variables
Turbine
Q
(hf , d, g, L, ν) into a single dimensionless group, with hf
3
2
d in the numerator, denoted as ξ, which equals (f Red /2).
(b) Rearrange the Colebrook formula (6.48) to solve for
10 m
30 m
Red in terms of ξ. (c) For extra credit, solve Example 6.9
D = 6 cm
D = 4 cm
with this new formula.
P6.69 For Prob. P6.62 suppose the only pump available can
P6.76
­deliver 80 hp to the fluid. What is the proper pipe size in
3
inches to maintain the 3 ft /s flow rate?
*P6.77 Modify Prob. P6.76 into an economic analysis, as folP6.70 Ethylene glycol at 20°C flows through 80 m of cast iron
lows: Let the 40 m of wrought iron pipe have a uniform
pipe of diameter 6 cm. The measured pressure drop is
diameter d. Let the steady water flow available be Q =
250 kPa. Neglect minor losses. Using a noniterative for30 m3/h. The cost of the turbine is $4 per watt developed,
3
mulation, estimate the flow rate in m /h.
and the cost of the piping is $75 per centimeter of
440
Chapter 6 Viscous Flow in Ducts
d­ iameter. The power generated may be sold for $0.08 per
kilowatt-hour. Find the proper pipe diameter for minimum payback time—that is, the minimum time for
which the power sales will equal the initial cost of the
system.
P6.78 In Fig. P6.78 the connecting pipe is commercial steel 6
cm in diameter. Estimate the flow rate, in m3/h, if the
fluid is water at 20°C. Which way is the flow?
P6.79 A garden hose is to be used as the return line in a waterfall display at a mall. In order to select the proper
pump, you need to know the roughness height inside
the garden hose. Unfortunately, roughness information
is not supplied by the hose manufacturer. So you devise
a simple experiment to measure the roughness. The
hose is attached to the drain of an above-ground swimming pool, the surface of which is 3.0 m above the hose
outlet. You estimate the minor loss coefficient of the
entrance region as 0.5, and the drain valve has a minor
loss equivalent length of 200 diameters when fully
open. Using a bucket and stopwatch, you open the valve
and measure the flow rate to be 2.0 × 10−4 m3/s for a
hose that is 10.0 m long and has an inside diameter of
1.50 cm. Estimate the roughness height in mm inside
the hose.
200 kPa
gage
15 m
L = 50 m
P6.78
P6.80 The head-versus-flow-rate characteristics of a centrifugal pump are shown in Fig. P6.80. If this pump
drives water at 20°C through 120 m of 30-cm-diameter cast iron pipe, what will be the resulting flow rate,
in m3/s?
P6.81 The pump in Fig. P6.80 is used to deliver gasoline at
20°C through 350 m of 30-cm-diameter galvanized iron
pipe. Estimate the resulting flow rate, in m3/s. (Note that
the pump head is now in meters of gasoline.)
P6.82 Fluid at 20°C flows through a horizontal galvanized-iron
pipe 20 m long and 8 cm in diameter. The wall shear
stress is 90 Pa. Calculate the flow rate in m3/h if the fluid
is (a) glycerin and (b) water.
P6.83 For the system of Fig. P6.55, let Δz = 80 m and L = 185
m of cast iron pipe. What is the pipe diameter for which
the flow rate will be 7 m3/h?
P6.84 It is desired to deliver 60 m3/h of water at 20°C through
a horizontal asphalted cast iron pipe. Estimate the pipe
diameter that will cause the pressure drop to be exactly
40 kPa per 100 m of pipe length.
P6.85 For the system in Prob. P6.53, a pump, which delivers
15,000 W to the water, is used at night to refill the upper
reservoir. The pipe diameter is increased from 12 cm to
provide more flow. If the resultant flow rate is 90 m3/h,
estimate the new pipe size.
Noncircular ducts
P6.86 SAE 10 oil at 20°C flows at an average velocity of 2
m/s between two smooth parallel horizontal plates 3
cm apart. Estimate (a) the centerline velocity, (b) the
head loss per meter, and (c) the pressure drop per
meter.
P6.87 A commercial steel annulus 40 ft long, with a = 1 in and
b = 12 in, connects two reservoirs that differ in surface
height by 20 ft. Compute the flow rate in ft3/s through
the annulus if the fluid is water at 20°C.
P6.88 An oil cooler consists of multiple parallel-plate passages, as shown in Fig. P6.88. The available pressure
drop is 6 kPa, and the fluid is SAE 10W oil at 20°C.
If the desired total flow rate is 900 m3/h, estimate the
­appropriate number of passages. The plate walls are
­hydraulically smooth.
2m
80 m
hp
Parabola
Flow
Pump
performance
P6.80
0
Q
2 m3/s
P6.88
50 cm
50 cm
Problems 441
P6.89 An annulus of narrow clearance causes a very large pressure drop and is useful as an accurate measurement of
viscosity. If a smooth annulus 1 m long with a = 50 mm
and b = 49 mm carries an oil flow at 0.001 m3/s, what is
the oil viscosity if the pressure drop is 250 kPa?
P6.90 A rectangular sheet-metal duct is 200 ft long and has a
fixed height H = 6 in. The width B, however, may vary
from 6 to 36 in. A blower provides a pressure drop of 80
Pa of air at 20°C and 1 atm. What is the optimum width
B that will provide the most airflow in ft3/s?
P6.91 Heat exchangers often consist of many triangular passages. Typical is Fig. P6.91, with L = 60 cm and an
isosceles-triangle cross section of side length a = 2 cm
and included angle β = 80°. If the average velocity is
V = 2 m/s and the fluid is SAE 10 oil at 20°C, estimate
the pressure drop.
a
β
P6.94 Air at 20°C flows through a smooth duct of d­ iameter 20
cm at an average velocity of 5 m/s. It then flows into a
smooth square duct of side length a. Find the square duct
size a for which the pressure drop per meter will be exactly the same as the circular duct.
P6.95 Although analytical solutions are available for laminar
flow in many duct shapes [34], what do we do about
ducts of arbitrary shape? Bahrami et al. [57] propose that
a better approach to the pipe result, f Re = 64, is achieved
by replacing the hydraulic diameter Dh with √A, where
A is the area of the cross section. Test this idea for the
isosceles triangles of Table 6.4. If time is short, at least
try 10°, 50°, and 80°. What do you conclude about this
idea?
P6.96 A fuel cell [45] consists of air (or oxygen) and hydrogen
micro ducts, separated by a membrane that promotes
proton exchange for an electric current, as in Fig. P6.96.
Suppose that the air side, at 20°C and approximately 1
atm, has five 1 mm by 1 mm ducts, each 1 m long. The
total flow rate is 1.5 E−4 kg/s. (a) Determine if the flow
is laminar or turbulent. (b) Estimate the pressure drop.
(Problem courtesy of Dr. Pezhman Shirvanian.)
L
V
Air
flow
Hydrogen
flow
P6.91
P6.92 A large room uses a fan to draw in atmospheric air at
20°C through a 30-cm by 30-cm commercial-steel duct
12 m long, as in Fig. P6.92. Estimate (a) the airflow rate
in m3/h if the room pressure is 10 Pa vacuum and (b) the
room pressure if the flow rate is 1200 m3/h. Neglect minor losses.
Fan
Room
30 cm by 30 cm
patm
12 m
P6.92
P6.93 In Moody’s Example 6.6, the 6-inch diameter, 200-ftlong asphalted cast iron pipe has a pressure drop of about
280 lbf/ft2 when the average water velocity is 6 ft/s.
Compare this to an annular cast iron pipe with an inner
diameter of 6 in and the same annular average velocity of
6 ft/s. (a) What outer diameter would cause the flow to
have the same pressure drop of 280 lbf/ft2? (b) How do
the cross-section areas compare, and why? Use the hydraulic diameter approximation.
P6.96
Anode
Cathode
1 mm by 1 mm by 1 m
PEM membrane
P6.97 A heat exchanger consists of multiple parallel-plate passages, as shown in Fig. P6.97. The available pressure
drop is 2 kPa, and the fluid is water at 20°C. If the desired
total flow rate is 900 m3/h, estimate the appropriate number of passages. The plate walls are hydraulically smooth.
442
Chapter 6 Viscous Flow in Ducts
2m
50 cm
Flow
*P6.100 Modify Prob. P6.55 as follows: Assume a pump can deliver 3 kW to pump the water back up to reservoir 1 from
reservoir 2. Accounting for an open flanged globe valve
and sharp-edged entrance and exit, estimate the predicted flow rate, in m3/h.
50 cm P6.101 In Fig. P6.101 a thick filter 3is being tested for losses. The
flow rate in the pipe is 7 m /min, and the upstream pressure is 120 kPa. The fluid is air at 20°C. Using the water
manometer reading, estimate the loss coefficient K of
the filter.
P6.97
Air
P6.98 A rectangular heat exchanger is to be divided into
smaller sections using sheets of commercial steel
0.4 mm thick, as sketched in Fig. P6.98. The flow rate
is 20 kg/s of water at 20°C. Basic dimensions are L =
1 m, W = 20 cm, and H = 10 cm. What is the proper
number of square sections if the overall pressure drop
is to be no more than 1600 Pa?
W
d = 10 cm
4 cm
Water
P6.101
*P6.102 A 70 percent efficient pump delivers water at 20°C from
one reservoir to another 20 ft higher, as in Fig. P6.102.
The piping system consists of 60 ft of galvanized iron
2-in pipe, a reentrant entrance, two screwed 90° longradius elbows, a screwed-open gate valve, and a sharp
exit. What is the input power required in horsepower
with and without a 6° well-designed conical expansion
added to the exit? The flow rate is 0.4 ft3/s.
H
L
P6.98
Minor or local losses
P6.99 In Sec. 6.11 it was mentioned that Roman aqueduct customers obtained extra water by attaching a diffuser to
their pipe exits. Figure P6.99 shows a simulation: a
smooth inlet pipe, with or without a 15° conical diffuser
expanding to a 5-cm-diameter exit. The pipe entrance is
sharp-edged. Calculate the flow rate (a) without and (b)
with the diffuser.
6° cone
20 ft
Pump
D2 = 5 cm
2m
D1 = 3 cm, L = 2 m
15° diffuser
P6.99
P6.102
P6.103 The reservoirs in Fig. P6.103 are connected by cast iron
pipes joined abruptly, with sharp-edged entrance and
exit. Including minor losses, estimate the flow of water
at 20°C if the surface of reservoir 1 is 45 ft higher than
that of reservoir 2.
Problems 443
*P6.107 A tank of water 4 m in diameter and 7 m deep is to be
drained by a 5-cm-diameter exit pipe at the bottom, as in
Fig. P6.107. In design (1), the pipe extends out for 1 m and
into the tank for 10 cm. In design (2), the interior pipe is
removed and the entrance beveled, Fig. 6.21, so that K ≈ 0.1
in the entrance. (a) An engineer claims that design (2) will
drain 25 percent faster than design (1). Is this claim true? (b)
Estimate the time to drain of design (2), assuming f ≈ 0.020.
D = 2 in
L = 20 ft
1
1 in
2
2 in
D = 1 in
L = 20 ft
P6.103
P6.104 Consider a 20°C flow at 2 m/s through a smooth 3-mm
­diameter microtube which consists of a straight run of
10 cm, a long radius bend, and another straight run of
10 cm. Compute the total pressure drop if the fluid is
(a) water and (b) ethylene glycol.
P6.105 The system in Fig. P6.105 consists of 1200 m of 5 cm
cast iron pipe, two 45° and four 90° flanged long-radius
elbows, a fully open flanged globe valve, and a sharp
exit into a reservoir. If the elevation at point 1 is 400 m,
what gage pressure is required at point 1 to deliver 0.005
m3/s of water at 20°C into the reservoir?
(1)
(2)
P6.107
P6.108 The water pump in Fig. P6.108 maintains a pressure of
6.5 psig at point 1. There is a filter, a half-open disk
valve, and two regular screwed elbows. There are 80 ft of
4-in diameter commercial steel pipe. (a) If the flow rate
is 0.4 ft3/s, what is the loss coefficient of the filter? (b) If
the disk valve is wide open and Kfilter = 7, what is the
resulting flow rate?
Elevation
500 m
9 ft
Sharp
exit
45°
1
1
Filter
Valve
Elbows
Open
globe
Pump
P6.108
45°
P6.105
P6.106 The water pipe in Fig. P6.106 slopes upward at 30°. The
pipe has a 1-in diameter and is smooth. The flanged
globe valve is fully open. If the mercury manometer
shows a 7-in deflection, what is the flow rate in ft3/s?
P6.109 In Fig. P6.109 there are 125 ft of 2-in pipe, 75 ft of 6-in
pipe, and 150 ft of 3-in pipe, all cast iron. There are three
90° elbows and an open globe valve, all flanged. If the
exit elevation is zero, what horsepower is extracted by the
turbine when the flow rate is 0.16 ft3/s of water at 20°C?
Elevation 100 ft
2 in
Globe
Turbine
7 in
Mercury
6 in
10 ft
P6.106
P6.109
Open
globe
3 in
444
Chapter 6 Viscous Flow in Ducts
P6.110 In Fig. P6.110 the pipe entrance is sharp-edged. If the flow *P6.114 A blower supplies standard air to a plenum that feeds
rate is 0.004 m3/s, what power, in W, is extracted by the
two horizontal square sheet-metal ducts with sharpturbine?
edged entrances. One duct is 100 ft long, with a crosssection 6 in by 6 in. The second duct is 200 ft long. Each
duct exhausts to the atmosphere. When the plenum
Open globe
Turbine
valve
pressure is 5.0 lbf/ft2 (gage) the volume flow in the lon40 m
ger duct is three times the flow in the shorter duct. Estimate both volume flows and the cross-section size of
the longer duct.
Water
P6.115 In Fig. P6.115 all pipes are 8-cm-diameter cast iron. DeCast iron:
termine the flow rate from reservoir 1 if valve C is (a)
L = 125 m, D = 5 cm
closed and (b) open, K = 0.5.
P6.110
1 Z = 25 m
Series and parallel pipe systems
P6.111 For the parallel-pipe system of Fig. P6.111, each pipe is
cast iron, and the pressure drop p1 − p2 = 3 lbf/in2. Compute the total flow rate between 1 and 2 if the fluid is
SAE 10 oil at 20°C.
D = 3 in
2 Z=0m
Water
at 20°C
C
A
30 m
L = 70 m
L = 200 ft
1
10 m
B
L = 100 m
L = 250 ft
D = 2 in
L = 50 m
Valve
2
P6.111
P6.115
P6.112 If the two pipes in Fig. P6.111 are instead laid in series
with the same total pressure drop of 3 lbf/in2, what will
the flow rate be? The fluid is SAE 10 oil at 20°C.
P6.113 The parallel galvanized iron pipe system of Fig. P6.113
delivers water at 20°C with a total flow rate of 0.036
m3/s. If the pump is wide open and not running, with a
loss coefficient K = 1.5, determine (a) the flow rate in
each pipe and (b) the overall pressure drop.
P6.116 For the series-parallel system of Fig. P6.116, all pipes
are 8-cm-diameter asphalted cast iron. If the total
pressure drop p1 − p2 = 750 kPa, find the resulting
flow rate Q m3/h for water at 20°C. Neglect minor
losses.
L = 250 m
L1 = 60 m, D1 = 5 cm
150 m
100 m
Pump
L 2 = 55 m, D2 = 4 cm
P6.113
Q = 0.036 m3/s
P6.116
1
2
P6.117 A blower delivers air at 3000 m3/h to the duct circuit in
Fig. P6.117. Each duct is commercial steel and of square
cross section, with side lengths a1 = a3 = 20 cm and
a2 = a4 = 12 cm. Assuming sea-level air conditions, estimate the power required if the blower has an efficiency
of 75 percent. Neglect minor losses.
Problems 445
Three-reservoir and pipe network systems
P6.121 Consider the three-reservoir system of Fig. P6.121 with
the following data:
3
4
2
z1 = 25 m
1
Blower
L1 = 95 m L2 = 125 m L3 = 160 m
30 m
Z2
Z3
Z1
P6.118 For the piping system of Fig. P6.118, all pipes are concrete with a roughness of 0.04 in. Neglecting minor
losses, compute the overall pressure drop p1 − p2 in lbf/
in2 if Q =20 ft3/s. The fluid is water at 20°C.
D = 8 in
L = 1500 ft
D = 12 in
D = 12 in
L = 1000 ft
L = 800 ft
2
P6.118
P6.119 For the piping system of Prob. P6.111, let the fluid be
gasoline at 20°C, with both pipes cast iron. If the flow
rate in the 2-in pipe (b) is 1.2 ft3/min, estimate the flow
rate in the 3-in pipe (a), in ft3/min.
P6.120 Three cast iron pipes are laid in parallel with these
­dimensions:
Pipe
Length, m
1
2
3
800
600
900
L2
L3
L1
P6.121
D = 15 in
L = 1200 ft
z3 = 85 m
All pipes are 28-cm-diameter unfinished concrete
(ε = 1 mm). Compute the steady flow rate in all pipes
for water at 20°C.
40 m
P6.117
1
z2 = 115 m
P6.122 Modify Prob. P6.121 as follows: Reduce the diameter to
15 cm (with ε = 1 mm), and compute the flow rates for
water at 20°C. These flow rates distribute in nearly the
same manner as in Prob. P6.121 but are about 5.2 times
lower. Can you explain this difference?
P6.123 Modify Prob. P6.121 as follows: All data are the same
except that z3 is unknown. Find the value of z3 for
which the flow rate in pipe 3 is 0.2 m3/s toward the
junction. (This problem requires iteration and is best
suited to a computer.)
P6.124 The three-reservoir system in Fig. P6.124 delivers water
at 20°C. The system data are as follows:
D1 = 8 in
D2 = 6 in
D3 = 9 in
L1 = 1800 ft
L2 = 1200 ft
L3 = 1600 ft
All pipes are galvanized iron. Compute the flow rate in
all pipes.
z2 = 100 ft
Diameter, cm
12
8
10
The total flow rate is 200 m3/h of water at 20°C. Determine (a) the flow rate in each pipe and (b) the pressure
drop across the system.
z3 = 50 ft
z1 = 20 ft
2
3
1
P6.124
J
446
Chapter 6 Viscous Flow in Ducts
P6.125 Suppose that the three cast iron pipes in Prob. P6.120 are
instead connected to meet smoothly at a point B, as
shown in Fig. P6.125. The inlet pressures in each pipe
are
p1 = 200 kPa p2 = 160 kPa p3 = 100 kPa.
The fluid is water at 20°C. Neglect minor losses. Estimate the flow rate in each pipe and whether it is toward
or away from point B.
1
P6.128 Modify Prob. P6.127 as follows: Let the inlet flow
rate at A and the exit flow at D be unknown. Let pA −
pB = 100 lbf/in2. Compute the flow rate in all five
pipes.
P6.129 In Fig. P6.129 all four horizontal cast iron pipes are 45 m
long and 8 cm in diameter and meet at junction a, delivering water at 20°C. The pressures are known at four
points as shown:
p1 = 950 kPa p2 = 350 kPa
2
p3 = 675 kPa p4 = 100 kPa
Neglecting minor losses, determine the flow rate in each
pipe.
B
p1
p2
L1
L2
3
a
P6.125
P6.126 Modify Prob. P6.124 as follows: Let all data be the same
except that pipe 1 is fitted with a butterfly valve
(Fig. 6.19b). Estimate the proper valve opening angle (in
degrees) for the flow rate through pipe 1 to be reduced to
1.5 ft3/s toward reservoir 1. (This problem requires iteration and is best suited to a computer.)
P6.127 In the five-pipe horizontal network of Fig. P6.127,
assume that all pipes have a friction factor f = 0.025.
For the given inlet and exit flow rate of 2 ft3/s of water
at 20°C, determine the flow rate and direction in all
pipes. If pA = 120 lbf/in2 gage, determine the pressures at points B, C, and D.
d = 8 in
2 ft3/s
D
C
p4
P6.130 In Fig. P6.130 lengths AB and BD are 2000 and
1500 ft, respectively. The friction factor is 0.022 everywhere, and pA = 90 lbf/in2 gage. All pipes have a
­diameter of 6 in. For water at 20°C, determine the
flow rate in all pipes and the pressures at points B, C,
and D.
0.5 ft3/s
2 ft3/s
P6.127
3000 ft
A
8 in
4000 ft
0.5 ft3/s
D
C
3 in
A
p3
L4
P6.129
9 in
6 in
L3
B
2.0 ft3/s
P6.130
B
1.0 ft3/s
Problems 447
Diffuser performance
P6.131 A water tunnel test section has a 1-m diameter and flow
properties V = 0.2 m/s, p = 100 kPa, and T = 20°C. The
boundary layer blockage at the end of the section is
9 ­percent. If a conical diffuser is to be added at the end of
the section to achieve maximum pressure recovery, what
should its angle, length, exit diameter, and exit pressure
be?
P6.132 For Prob. P6.131 suppose we are limited by space to a
total diffuser length of 10 m. What should the diffuser
angle, exit diameter, and exit pressure be for maximum
recovery?
P6.133 A wind tunnel test section is 3 ft square with flow properties V = 15 ft/s, p = 15 lbf/in2 absolute, and T = 68°F.
Boundary layer blockage at the end of the test section is
8 percent. Find the angle, length, exit height, and exit
pressure of a flat-walled diffuser added onto the section
to achieve maximum pressure recovery.
P6.134 For Prob. P6.133 suppose we are limited by space to a
total diffuser length of 30 ft. What should the diffuser
angle, exit height, and exit pressure be for maximum recovery?
The pitot-static tube
P6.135 An airplane uses a pitot-static tube as a velocimeter. The
measurements, with their uncertainties, are a static temperature of (−11 ± 3)°C, a static pressure of 60 ± 2 kPa,
and a pressure difference (po − ps) = 3200 ± 60 Pa.
(a) ­Estimate the airplane’s velocity and its uncertainty.
(b) Is a compressibility correction needed?
P6.136 For the pitot-static pressure arrangement of Fig. P6.136,
the manometer fluid is (colored) water at 20°C. Estimate
(a) the centerline velocity, (b) the pipe volume flow, and
(c) the (smooth) wall shear stress.
Air
8 cm
20°C
1 atm
40 mm
P6.136
P6.137 For the 20°C water flow of Fig. P6.137, use the pitotstatic arrangement to estimate (a) the centerline velocity
and (b) the volume flow in the 5-in-diameter smooth
pipe. (c) What error in flow rate is caused by neglecting
the 1-ft elevation difference?
1 ft
2 in
Mercury
P6.137
P6.138 An engineer who took college fluid mechanics on a
pass–fail basis has placed the static pressure hole far
upstream of the stagnation probe, as in Fig. P6.138,
thus contaminating the pitot measurement ridiculously with pipe friction losses. If the pipe flow is air
at 20°C and 1 atm and the manometer fluid is Meriam
red oil (SG = 0.827), estimate the air centerline velocity for the given manometer reading of 16 cm. Assume a smooth-walled tube.
10 m
Air
16 cm
D = 6 cm
P6.138
P6.139 Professor Walter Tunnel needs to measure the flow
­velocity in a water tunnel. Due to budgetary restrictions,
he cannot afford a pitot-static probe, but instead inserts a
total head probe and a static pressure probe, as shown in
Fig. P6.139, a distance h1 apart from each other. Both
probes are in the main free stream of the water tunnel,
­unaffected by the thin boundary layers on the sidewalls.
The two probes are connected as shown to a U-tube
­manometer. The densities and vertical distances are
shown in Fig. P6.139. (a) Write an expression for velocity V in terms of the parameters in the problem. (b) Is it
critical that distance h1 be measured accurately? (c) How
does the ­expression for velocity V differ from that which
would be obtained if a pitot-static probe had been available and used with the same U-tube manometer?
448
Chapter 6 Viscous Flow in Ducts
ptotal
h1
V
h2
pstatic
ρw
h3
P6.143 A 10-cm-diameter smooth pipe contains an orifice plate
with D: 12 D taps and β = 0.5. The measured orifice pressure drop is 75 kPa for water flow at 20°C. Estimate the
flow rate, in m3/h. What is the nonrecoverable head loss?
*P6.144 Water at 20°C flows through the orifice in Fig. P6.154,
which is monitored by a mercury manometer. If d = 3 cm,
(a) what is h when the flow rate is 20 m3/h and (b) what
is Q in m3/h when h = 58 cm?
5 cm
Water
d
U-tube manometer
h
ρm
P6.139
Mercury
P6.144
Flowmeters: the orifice plate
P6.140 Gasoline at 20°C flows at 3 m3/h in a 6-cm-diameter
pipe. A 4-cm-diameter thin-plate orifice with corner taps
is ­installed. Estimate the measured pressure drop, in Pa.
P6.141 Gasoline at 20°C flows at 105 m3/h in a 10-cm-diameter pipe. We wish to meter the flow with a thin-plate
orifice and a differential pressure transducer that
reads best at about 55 kPa. What is the proper β ratio
for the orifice?
P6.142 The shower head in Fig. P6.142 delivers water at 50°C.
An orifice-type flow reducer is to be installed. The upstream pressure is constant at 400 kPa. What flow rate,
in gal/min, results without the reducer? What reducer
orifice diameter would decrease the flow by 40 percent?
D = 1.5 cm
p = 400 kPa
P6.145 The 1-m-diameter tank in Fig. P6.145 is initially filled
with gasoline at 20°C. There is a 2-cm-diameter orifice
in the bottom. If the orifice is suddenly opened, estimate
the time for the fluid level h(t) to drop from 2.0 to 1.6 m.
h (0) = 2 m
1m
P6.145
h (t )
Q (t )
P6.146 A pipe connecting two reservoirs, as in Fig. P6.146, contains a thin-plate orifice. For water flow at 20°C, estimate (a) the volume flow through the pipe and (b) the
pressure drop across the orifice plate.
Flow reducer
20 m
45 holes, 1.5-mm diameter
P6.142
P6.146
L = 100 m
D = 5 cm
3-cm
orifice
Problems 449
P6.147 Air flows through a 6-cm-diameter smooth pipe that has
a 2-m-long perforated section containing 500 holes (diameter 1 mm), as in Fig. P6.147. Pressure outside the
pipe is sea-level standard air. If p1 = 105 kPa and Q1 =
110 m3/h, estimate p2 and Q2, assuming that the holes are
approximated by thin-plate orifices. (Hint: A momentum control volume may be very useful.)
versus tank pressure. Is the flow laminar or turbulent?
Compare the data with theoretical results obtained from
the Moody chart, including minor losses. Discuss.
pa = 1 atm
Air
tank
8m
pgage
500 holes (diameter 1 mm)
V
h
1
2
D = 6 cm
2m
P6.149
P6.147
P6.148 A smooth pipe containing ethanol at 20°C flows at
7 m3/h through a Bernoulli obstruction, as in Fig. P6.148.
Three piezometer tubes are installed, as shown. If the
obstruction is a thin-plate orifice, estimate the piezometer levels (a) h2 and (b) h3.
h3
h2
h 1= 1 m
5m
d = 3 cm
D = 5 cm
P6.148
P6.150 Gasoline at 20°C flows at 0.06 m3/s through a 15-cm
pipe and is metered by a 9-cm long-radius flow nozzle
(Fig. 6.40a). What is the expected pressure drop across
the nozzle?
P6.151 An engineer needs to monitor a flow of 20°C gasoline at
about 250 ± 25 gal/min through a 4-in-diameter smooth
pipe. She can use an orifice plate, a long-radius flow
nozzle, or a venturi nozzle, all with 2-in-diameter
throats. The only differential pressure gage available is
accurate in the range 6 to 10 lbf/in2. Disregarding flow
losses, which ­device is best?
P6.152 Kerosene at 20°C flows at 20 m3/h in an 8-cm-diameter
pipe. The flow is to be metered by an ISA 1932 flow
nozzle so that the pressure drop is 7000 Pa. What is the
proper nozzle diameter?
P6.153 Two water tanks, each with base area of 1 ft2, are
­connected by a 0.5-in-diameter long-radius nozzle as in
Fig. P6.153. If h = 1 ft as shown for t = 0, estimate the
time for h(t) to drop to 0.25 ft.
Flowmeters: the flow nozzle
h = 1 ft
P6.149 In a laboratory experiment, air at 20°C flows from a
large tank through a 2-cm-diameter smooth pipe into a
sea-level atmosphere, as in Fig. P6.149. The flow is metered by a long-radius nozzle of 1-cm diameter, using a
manometer with Meriam red oil (SG = 0.827). The pipe
is 8 m long. The measurements of tank pressure and manometer height are as follows:
ptank, Pa(gage):
60
320
1200
2050
2470
3500
4900
hmano, mm:
6
38
160
295
380
575
820
Use these data to calculate the flow rates Q and Reynolds numbers ReD and make a plot of measured flow rate
d = 12 in
P6.153
1 ft 2
2 ft
1 ft 2
Flowmeters: the venturi meter
P6.154 Gasoline at 20°C flows through a 6-cm-diameter pipe. It
is metered by a modern venturi nozzle with d = 4 cm.
450
Chapter 6 Viscous Flow in Ducts
The measured pressure drop is 8.5 kPa. Estimate the
flow rate in gallons per minute.
P6.155 It is desired to meter methanol at 20°C flowing
through a 5-in-diameter pipe. The expected flow rate
is about 300 gal/min. Two flowmeters are available: a
venturi nozzle and a thinplate orifice, each with d = 2
in. The differential pressure gage on hand is most accurate at about 12–15 lbs/in2. Which meter is better
for this job?
P6.156 Ethanol at 20°C flows down through a modern venturi
nozzle as in Fig. P6.156. If the mercury manometer
reading is 4 in, as shown, estimate the flow rate, in
gal/min.
Use these data to plot a calibration curve of venturi discharge coefficient versus Reynolds number. Compare
with the accepted correlation, Eq. (6.115).
Flowmeters: other designs
P6.160 An instrument popular in the beverage industry is the
target flowmeter in Fig. P6.160. A small flat disk is
mounted in the center of the pipe, supported by a strong
but thin rod. (a) Explain how the flowmeter works. (b) If
the bending moment M of the rod is measured at the
wall, derive a formula for the estimated velocity of the
flow. (c) List a few advantages and disadvantages of
such an instrument.
D = 6 in
9 in
Flow
d = 3 in
4 in
P6.160
P6.156
P6.157 Modify Prob. P6.156 if the fluid is air at 20°C, entering
the venturi at a pressure of 18 lbf/in2. Should a compressibility correction be used?
P6.158 Water at 20°C flows in a long horizontal commercial
steel 6-cm-diameter pipe that contains a classical Herschel ­venturi with a 4-cm throat. The venturi is connected to a mercury manometer whose reading is
h = 40 cm. Estimate (a) the flow rate, in m3/h, and (b)
the total pressure ­difference between points 50 cm upstream and 50 cm downstream of the venturi.
P6.159 A modern venturi nozzle is tested in a laboratory flow
with water at 20°C. The pipe diameter is 5.5 cm, and the
venturi throat diameter is 3.5 cm. The flow rate is measured by a weigh tank and the pressure drop by a water–
mercury ­manometer. The mass flow rate and manometer
readings are as follows:

m , kg/s
0.95
1.98
2.99
5.06
8.15
h, mm
3.7
15.9
36.2
102.4
264.4
P6.161 An instrument popular in the water supply industry,
sketched in Fig. P6.161, is the single jet water meter.
(a) How does it work? (b) What do you think a typical
calibration curve would look like? (c) Can you cite
further details, for example, reliability, head loss, cost
[58]?
P6.161
Fundamentals of Engineering Exam Problems 451
Flowmeters: compressibility correction
D = 6 cm
d = 4 cm
P6.162 Air flows at high speed through a Herschel venturi monitored by a mercury manometer, as shown in Fig. P6.162.
The upstream conditions are 150 kPa and 80°C. If
h = 37 cm, estimate the mass flow in kg/s. (Hint: The
flow is compressible.)
P6.163 Modify Prob. P6.162 as follows: Find the manometer
­reading h for which the mass flow through the venturi is
approximately 0.4 kg/s. (Hint: The flow is compressible.)
Air
h
Mercury
P6.162
Word Problems
W6.1 In fully developed straight-duct flow, the velocity profiles do not change (why?), but the pressure drops along
the pipe axis. Thus there is pressure work done on the
fluid. If, say, the pipe is insulated from heat loss, where
does this energy go? Make a thermodynamic analysis of
the pipe flow.
W6.2 From the Moody chart (Fig. 6.13), rough surfaces, such
as sand grains or ragged machining, do not affect laminar
flow. Can you explain why? They do affect turbulent
flow. Can you develop, or suggest, an analytical–physical
model of turbulent flow near a rough surface that might
be used to predict the known increase in pressure drop?
W6.3 Differentiation of the laminar pipe flow solution, Eq.
(6.40), shows that the fluid shear stress τ(r) varies linearly
from zero at the axis to τw at the wall. It is claimed that this
is also true, at least in the time mean, for fully developed
turbulent flow. Can you verify this claim analytically?
W6.4 A porous medium consists of many tiny tortuous passages,
and Reynolds numbers based on pore size are usually very
low, of order unity. In 1856 H. Darcy proposed that the
pressure gradient in a porous medium was directly proportional to the volume-averaged velocity V of the fluid:
∇p = −
μ
V
K
where K is termed the permeability of the medium. This
is now called Darcy’s law of porous flow. Can you make
a Poiseuille flow model of porous-media flow that verifies Darcy’s law? Meanwhile, as the Reynolds number
increases, so that VK1/2/ν > 1, the pressure drop becomes
nonlinear, as was shown experimentally by P. H. Forscheimer as early as 1782. The flow is still decidedly
laminar, yet the pressure gradient is quadratic:
∇p = −
μ
V − C∣V∣V
K
Darcy−Forscheimer law
where C is an empirical constant. Can you explain the
­reason for this nonlinear behavior?
Fundamentals of Engineering Exam Problems
FE6.1 In flow through a straight, smooth pipe, the diameter
Reynolds number for transition to turbulence is generally
taken to be
(a) 1500, (b) 2300, (c) 4000, (d) 250,000, (e) 500,000
FE6.2 For flow of water at 20°C through a straight, smooth
pipe at 0.06 m3/h, the pipe diameter for which transition
to turbulence occurs is approximately
(a) 1.0 cm, (b) 1.5 cm, (c) 2.0 cm, (d) 2.5 cm, (e) 3.0 cm
FE6.3 For flow of oil [µ = 0.1 kg/(m · s), SG = 0.9] through a
long, straight, smooth 5-cm-diameter pipe at 14 m3/h,
the pressure drop per meter is approximately
(a) 2200 Pa, (b) 2500 Pa, (c) 10,000 Pa, (d) 160 Pa,
(e) 2800 Pa
FE6.4 For flow of water at a Reynolds number of 1.03
E6 through a 5-cm-diameter pipe of roughness
height 0.5 mm, the ­
a pproximate Moody friction
factor is
(a) 0.012, (b) 0.018, (c) 0.038, (d) 0.049, (e) 0.102
FE6.5 Minor losses through valves, fittings, bends, contractions, and the like are commonly modeled as proportional to (a) total head, (b) static head, (c) velocity head,
(d) ­pressure drop, (e) velocity
FE6.6 A smooth 8-cm-diameter pipe, 200 m long, connects
two reservoirs, containing water at 20°C, one of
which has a surface elevation of 700 m and the
other a surface elevation of 560 m. If minor losses
452
Chapter 6 Viscous Flow in Ducts
are neglected, the expected flow rate through the
pipe is
(a) 0.048 m3/h, (b) 2.87 m3/h, (c) 134 m3/h, (d) 172
m3/h, (e) 385 m3/h
FE6.7 If, in Prob. FE6.6 the pipe is rough and the actual flow
rate is 90 m3/h, then the expected average roughness
height of the pipe is approximately
(a) 1.0 mm, (b) 1.25 mm, (c) 1.5 mm, (d) 1.75 mm,
(e) 2.0 mm
FE6.8 Suppose in Prob. FE6.6 the two reservoirs are ­connected,
not by a pipe, but by a sharp-edged orifice of diameter
8 cm. Then the expected flow rate is ­approximately
(a) 90 m3/h, (b) 579 m3/h, (c) 748 m3/h, (d) 949 m3/h,
(e) 1048 m3/h
FE6.9 Oil [µ = 0.1 kg/(m · s), SG = 0.9] flows through a 50-mlong smooth 8-cm-diameter pipe. The maximum pressure
drop for which laminar flow is expected is approximately
(a) 30 kPa, (b) 40 kPa, (c) 50 kPa, (d) 60 kPa, (e) 70 kPa
FE6.10 Air at 20°C and approximately 1 atm flows through a
smooth 30-cm-square duct at 1500 ft3/min. The expected
pressure drop per meter of duct length is
(a) 1.0 Pa, (b) 2.0 Pa, (c) 3.0 Pa, (d) 4.0 Pa, (e) 5.0 Pa
FE6.11 Water at 20°C flows at 3 m3/h through a sharp-edged
3-­cm-diameter orifice in a 6-cm-diameter pipe. Estimate
the expected pressure drop across the orifice.
(a) 440 Pa, (b) 680 Pa, (c) 875 Pa, (d) 1750 Pa, (e) 1870
Pa
FE6.12 Water flows through a straight 10-cm-diameter pipe at a
diameter Reynolds number of 250,000. If the pipe roughness is 0.06 mm, what is the approximate Moody friction
factor?
(a) 0.015, (b) 0.017, (c) 0.019, (d) 0.026, (e) 0.032
FE6.13 What is the hydraulic diameter of a rectangular air-­
ventilation duct whose cross section is 1 m by 25 cm?
(a) 25 cm, (b) 40 cm, (c) 50 cm, (d) 75 cm, (e) 100 cm
FE6.14 Water at 20°C flows through a pipe at 300 gal/min with
a friction head loss of 45 ft. What is the power required
to drive this flow?
(a) 0.16 kW, (b) 1.88 kW, (c) 2.54 kW, (d) 3.41 kW,
(e) 4.24 kW
FE6.15 Water at 20°C flows at 200 gal/min through a pipe 150 m
long and 8 cm in diameter. If the friction head loss is
12 m, what is the Moody friction factor?
(a) 0.010, (b) 0.015, (c) 0.020, (d) 0.025, (e) 0.030
Comprehensive Problems
C6.1
*C6.2
A pitot-static probe will be used to measure the velocity
distribution in a water tunnel at 20°C. The two pressure
lines from the probe will be connected to a U-tube manometer that uses a liquid of specific gravity 1.7. The
maximum velocity expected in the water tunnel is 2.3
m/s. Your job is to select an appropriate U-tube from a
manufacturer that supplies manometers of heights 8, 12,
16, 24, and 36 in. The cost increases significantly with
manometer height. Which of these should you purchase?
A pump delivers a steady flow of water (ρ, µ) from a large
tank to two other higher-elevation tanks, as shown in Fig.
C6.2. The same pipe of diameter d and roughness ε is used
throughout. All minor losses except through the valve are
neglected, and the partially closed valve has a loss coefficient Kvalve. Turbulent flow may be assumed with all kinetic energy flux correction coefficients equal to 1.06.
The pump net head H is a known function of QA and hence
also of VA = QA/Apipe; for example, H = a − bVA2 , where a
and b are constants. Subscript J refers to the junction point
at the tee where branch A splits into B and C. Pipe length
LC is much longer than LB. It is desired to predict the pressure at J, the three pipe velocities and friction factors, and
the pump head. Thus there are eight variables: H, VA, VB,
VC, fA, fB, fC, pJ. Write down the eight equations needed to
resolve this problem, but do not solve, since an elaborate
iteration procedure would be required.
3
Large tank
VC
2
Branch B, LB
1
Large tank
C6.2
Pump
VA
Branch A, LA
J
VB
Large tank
Branch C, LC
Valve
Comprehensive Problems 453
C6.3
A small water slide is to be installed inside a swimming *C6.4
pool. See Fig. C6.3. The slide manufacturer recommends
a continuous water flow rate Q of 1.39 × 10−3 m3/s
(about 22 gal/min) down the slide, to ensure that the customers do not burn their bottoms. A pump is to be installed under the slide, with a 5.00-m-long,
4.00-cm-diameter hose supplying swimming pool water
for the slide. The pump is 80 ­percent efficient and will
rest fully submerged 1.00 m below the water surface.
The roughness inside the hose is about 0.0080 cm. The
hose discharges the water at the top of the slide as a free
jet open to the atmosphere. The hose outlet is 4.00 m
above the water surface. For fully developed turbulent
pipe flow, the kinetic energy flux correction factor is
about 1.06. Ignore any minor losses here. Assume that ρ
= 998 kg/m3 and υ = 1.00 × 10−6 m2/s for this water.
Find the brake horsepower (that is, the actual shaft power
in watts) required to drive the pump.
Suppose you build a rural house where you need to run a
pipe to the nearest water supply, which is fortunately at
an elevation of about 1000 m above that of your house.
The pipe will be 6.0 km long (the distance to the water
supply), and the gage pressure at the water supply is
1000 kPa. You require a minimum of 3.0 gal/min of water when the end of your pipe is open to the atmosphere.
To minimize cost, you want to buy the smallest-diameter
pipe possible. The pipe you will use is extremely smooth.
(a) Find the total head loss from the pipe inlet to its exit.
Neglect any minor losses due to valves, elbows, entrance
lengths, and so on, since the length is so long here and
major losses dominate. Assume the outlet of the pipe is
open to the a­ tmosphere. (b) Which is more important in
this problem, the head loss due to elevation difference or
the head loss due to pressure drop in the pipe? (c) Find
the minimum required pipe diameter.
Q
Weee!
Tube
4.00 m
Ladder
Sliding
board
Pump
Water
1.00 m
C6.3
C6.5
Water at room temperature flows at the same volume flow
rate, Q = 9.4 × 10−4 m3/s, through two ducts, one a round
pipe and one an annulus. The cross-sectional area A of the
two ducts is identical, and all walls are made of commercial steel. Both ducts are the same length. In the cross sections shown in Fig. C6.5, R = 15.0 mm and a = 25.0 mm.
A
C6.6
b
a
R
C6.5
(a) What is the radius b such that the cross-sectional areas of the two ducts are identical? (b) Compare the frictional head loss hf per unit length of pipe for the two
cases, ­assuming fully developed flow. For the annulus,
do both a quick estimate (using the hydraulic diameter)
and a more accurate estimate (using the effective diameter correction), and compare. (c) If the losses are different for the two cases, explain why. Which duct, if any, is
more “efficient”?
John Laufer (NACA Tech Rep. 1174, 1954) gave velocity
data 20°C airflow in a smooth 24.7-cm-diameter pipe at
Re ≈ 5 E5:
u/uCL
1.0 0.997
0.988 0.959 0.908
0.847 0.818 0.771 0.690
r/R
0.0 0.102 0.206 0.412 0.617 0.784 0.846 0.907 0.963
Source: John Laufer (NASA Tech Rep. 1174, 1954)
The centerline velocity uCL was 30.5 m/s. Determine
(a) the average velocity by numerical integration and
454
C6.7
C6.8
Chapter 6 Viscous Flow in Ducts
(b) the wall shear stress from the log law approximation. C
­ ompare with the Moody chart and with
Eq. (6.43).
Consider energy exchange in fully developed laminar
flow between parallel plates, as in Eqs. (6.60). Let the
pressure drop over a length L be Δp. Calculate the rate
of work done by this pressure drop on the fluid in the
region (0 < x < L, −h < y < +h) and compare with the
integrated energy dissipated due to the viscous function
Φ from Eq. (4.50) over this same region. The two
should be equal. Explain why this is so. Can you relate
the viscous drag force and the wall shear stress to this
energy result?
This text has presented the traditional correlations for
the turbulent smooth-wall friction factor, Eq. (6.38),
and the law of the wall, Eq. (6.28). Recently, groups at
Princeton and Oregon [56] have made new friction
measurements and suggest the following smooth-wall
friction law:
1
√f
C6.9
In earlier work, they also report that better values for the
constants κ and B in the log-law, Eq. (6.28), are κ ≈
0.421 ± 0.002 and B ≈ 5.62 ± 0.08. (a) Calculate a few
values of f in the range 1 E4 ≤ ReD ≤ 1 E8 and see how
the two formulas differ. (b) Read Ref. 56 and briefly
check the five papers in its bibliography. Report to the
class on the general results of this work.
A pipeline has been proposed to carry natural gas
1715 miles from Alaska’s North Slope to Calgary, Alberta, Canada. The (assumed smooth) pipe diameter will
be 52 in. The gas will be at high pressure, averaging
2500 lbs/in2. (a) Why? The proposed flow rate is 4 billion cubic feet per day at sea-level conditions. (b) What
volume flow rate, at 20°C, would carry the same mass at
the high pressure? (c) If natural gas is assumed to be
methane (CH4), what is the total pressure drop? (d) If
each pumping station can deliver 12,000 hp to the flow,
how many stations are needed?
= 1.930 log10 (ReD √f ) − 0.537
Design Projects
D6.1
A hydroponic garden uses the 10-m-long perforatedpipe system in Fig. D6.1 to deliver water at 20°C. The
pipe is 5 cm in diameter and contains a circular hole
every 20 cm. A pump delivers water at 75 kPa (gage)
at the entrance, while the other end of the pipe is
closed. If you attempted, for example, Prob. P3.125,
you know that the pressure near the closed end of a
perforated “manifold” is surprisingly high, and there
will be too much flow through the holes near that end.
One remedy is to vary the hole size along the pipe axis.
Make a design analysis, perhaps using a personal computer, to pick the optimum hole size distribution that
will make the discharge flow rate as uniform as possible along the pipe axis. You are constrained to pick
hole sizes that correspond only to commercial (numbered) metric drill-bit sizes available to the typical machine shop.
20 cm
Pump
D6.1
10 m
References 455
D6.2
It is desired to design a pump-piping system to keep a
1-million-gallon capacity water tank filled. The plan is
to use a modified (in size and speed) version of the
model 1206 centrifugal pump manufactured by Taco
Inc., ­Cranston, Rhode Island. Test data have been provided to us by Taco Inc. for a small model of this
pump: D = 5.45 in, Ω = 1760 r/min, tested with water
at 20°C:
Q, gal/min
0
5
10 15 20 25
30 35
40 45
50 55 60
H, ft
28
28
29 29 28 28
27 26
25 23
21 18 15
Efficiency, % 0
13
25 35 44 48
51 53
54 55
53 50 45
The tank is to be filled daily with rather chilly (10°C)
groundwater from an aquifer, which is 0.8 mi from the
tank and 150 ft lower than the tank. Estimated daily water use is 1.5 million gal/day. Filling time should not
exceed 8 h per day. The piping system should have four
“butterfly” valves with variable openings (see Fig. 6.19),
10 elbows of various angles, and galvanized iron pipe of
a size to be selected in the design. The design should be
economical—both in capital costs and operating expense. Taco Inc. has provided the following cost estimates for system components:
Pump and motor$3500 plus $1500 per inch of impeller
size
Pump speed
Between 900 and 1800 r/min
Valves
$300 + $200 per inch of pipe size
Elbows$50 plus $50 per inch of pipe size
Pipes$1 per inch of diameter per foot
of length
Electricity cost 10¢ per kilowatt-hour
Your design task is to select an economical pipe size and
pump impeller size and speed for this task, using the
pump test data in nondimensional form (see Prob. P5.61)
as design data. Write a brief report (five to six pages)
showing your calculations and graphs.
References
1.
2.
3.
4.
5.
6.
7.
8.
9.
P. S. Bernard and J. M. Wallace, Turbulent Flow: Analysis, Measurement, and Prediction, Wiley, New York,
2002.
H. Schlichting et al., Boundary Layer Theory, Springer,
New York, 2000.
F. M. White, Viscous Fluid Flow, 3d ed., McGraw-Hill,
New York, 2005.
O. Reynolds, “An Experimental Investigation of the
Circumstances Which Determine Whether the Motion of
Water Shall Be Direct or Sinuous and of the Law of
Resistance in Parallel Channels,” Phil. Trans. R. Soc., vol.
174, 1883, pp. 935–982.
P. G. Drazin and W. H. Reid, Hydrodynamic Stability, 2d
ed., Cambridge University Press, New York, 2004.
H. Rouse and S. Ince, History of Hydraulics, Iowa Institute
of Hydraulic Research, State University of Iowa, Iowa City,
1957.
J. Nikuradse, “Strömungsgesetze in Rauhen Rohren,” VDI
Forschungsh. 361, 1933; English trans., NACA Tech.
Mem.1292.
L. F. Moody, “Friction Factors for Pipe Flow,” ASME
Trans., vol. 66, pp. 671–684, 1944.
C. F. Colebrook, “Turbulent Flow in Pipes, with Particular
Reference to the Transition between the Smooth and Rough
Pipe Laws,” J. Inst. Civ. Eng. Lond., vol. 11, 1938–1939,
pp. 133–156.
10.
11.
12.
13.
14.
15.
16.
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19.
O. C. Jones, Jr., “An Improvement in the Calculations of
Turbulent Friction in Rectangular Ducts,” J. Fluids Eng.,
June 1976, pp. 173–181.
R. Berker, Handbuch der Physik, vol. 7, no. 2, pp. 1–384,
Springer-Verlag, Berlin, 1963.
R. M. Olson, Essentials of Engineering Fluid Mechanics,
­Literary Licensing LLC, Whitefish, MT, 2012.
P. A. Durbin and B. A. Pettersson, Statistical Theory and
Modeling for Turbulent Flows, 2d ed., Wiley, New York,
2010.
P. W. Runstadler, Jr., et al., “Diffuser Data Book,” Creare
Inc. Tech. Note 186, Hanover, NH, 1975.
“Flow of Fluids through Valves, Fittings, and Pipes,” Tech.
­Paper 410, Crane Valve Group, Long Beach, CA, 1957
(now updated as a CD-ROM; see < http://www.cranevalves.com >).
E. F. Brater, H. W. King, J. E. Lindell, and C. Y. Wei,
Handbook of Hydraulics, 7th ed., McGraw-Hill, New
York, 1996.
H. Cross, “Analysis of Flow in Networks of Conduits or
Conductors,” Univ. Ill. Bull. 286, November 1936.
P. K. Swamee and A. K. Sharma, Design of Water Supply Pipe Networks, Wiley-Interscience, New York,
2008.
D. C. Wilcox, Turbulence Modeling for CFD, 3d ed., DCW
Industries, La Caňada, CA, 2006.
456
Chapter 6 Viscous Flow in Ducts
20.
R. W. Jeppson, Analysis of Flow in Pipe Networks,
­Butterworth-Heinemann, Woburn, MA, 1976.
R. W. Fox and S. J. Kline, “Flow Regime Data and Design
Methods for Curved Subsonic Diffusers,” J. Basic Eng.,
vol. 84, 1962, pp. 303–312.
R. C. Baker, Flow Measurement Handbook: Industrial
­Designs, Operating Principles, Performance, and Applications, Cambridge University Press, New York,
2005.
R. W. Miller, Flow Measurement Engineering Handbook,
3d edition, McGraw-Hill, New York, 1997.
B. Warren and C. Wunsch (eds.), Evolution of Physical
Oceanography, M.I.T. Press, Cambridge, MA, 1981.
U.S. Department of Commerce, Tidal Current Tables,
­National Oceanographic and Atmospheric Administration,
Washington, DC, 1971.
J. A. Shercliff, Electromagnetic Flow Measurement,
­Cambridge University Press, New York, 1962.
J. A. Miller, “A Simple Linearized Hot-Wire Anemometer,” J. Fluids Eng., December 1976, pp. 749–752.
R. J. Goldstein (ed.), Fluid Mechanics Measurements, 2d
ed., Hemisphere, New York, 1996.
D. Eckardt, “Detailed Flow Investigations within a High
Speed Centrifugal Compressor Impeller,” J. Fluids Eng.,
September 1976, pp. 390–402.
H. S. Bean (ed.), Fluid Meters: Their Theory and Application, 6th ed., American Society of Mechanical Engineers,
New York, 1971.
“Measurement of Fluid Flow by Means of Orifice Plates,
Nozzles, and Venturi Tubes Inserted in Circular Cross
­Section Conduits Running Full,” Int. Organ. Stand. Rep.
DIS-5167, Geneva, April 1976.
P. Sagaut and C. Meneveau, Large Eddy Simulation for
­Incompressible Flows: An Introduction, 3d ed., Springer,
New York, 2006.
S. E. Haaland, “Simple and Explicit Formulas for the Friction Factor in Turbulent Pipe Flow,” J. Fluids Eng., March
1983, pp. 89–90.
R. K. Shah and A. L. London, Laminar Flow Forced Convection in Ducts, Academic, New York, 1979.
P. L. Skousen, Valve Handbook, 3d ed. McGraw-Hill, New
York, 2011.
W. Li, W.-X. Chen, and S.-Z. Xie, “Numerical Simulation
of Stress-Induced Secondary Flows with Hybrid Finite
Analytic Method,” Journal of Hydrodynamics, vol. 14, no.
4, December 2002, pp. 24–30.
ASHRAE Handbook—2012 Fundamentals, ASHRAE,
­Atlanta, GA, 2012.
21.
22.
23.
24.
25.
26.
27.
28.
29.
30.
31.
32.
33.
34.
35.
36.
37.
38.
39.
40.
41.
42.
43.
44.
45.
46.
47.
48.
49.
50.
51.
52.
53.
54.
F. Durst, A. Melling, and J. H. Whitelaw, Principles and
Practice of Laser-Doppler Anemometry, 2d ed., Academic,
New York, 1981.
A. P. Lisitsyn et al., Laser Doppler and Phase Doppler
Measurement Techniques, Springer-Verlag, New York,
2003.
J. E. Amadi-Echendu, H. Zhu, and E. H. Higham, “Analysis of Signals from Vortex Flowmeters,” Flow Measurement and Instrumentation, vol. 4, no. 4, Oct. 1993,
pp. 225–231.
G. Vass, “Ultrasonic Flowmeter Basics,” Sensors, vol. 14,
no. 10, Oct. 1997, pp. 73–78.
ASME Fluid Meters Research Committee, “The ISOASME Orifice Coefficient Equation,” Mech. Eng. July
1981, pp. 44–45.
R. D. Blevins, Applied Fluid Dynamics Handbook, Van
­Nostrand Reinhold, New York, 1984.
O. C. Jones, Jr., and J. C. M. Leung, “An Improvement in
the Calculation of Turbulent Friction in Smooth C
­ oncentric
­Annuli,” J. Fluids Eng., December 1981, pp. 615–623.
C. Spiegel, Designing and Building Fuel Cells, McGrawHill, New York, 2007.
I. E. Idelchik, Handbook of Hydraulic Resistance, 3d ed.,
CRC Press, Boca Raton, FL, 1993.
B. A. Finlayson et al., Microcomponent Flow Characterization, Chap. 8 of Micro Instrumentation, M. V. Koch
(Ed.), John Wiley, Hoboken, NJ, 2007.
R. D. Coffield, P. T. McKeown, and R. B. Hammond,
“­Irrecoverable Pressure Loss Coefficients for Two Elbows
in Series with Various Orientation Angles and Separation
­Distances,” Report WAPD-T-3117, Bettis Atomic Power
Laboratory, West Mifflin, PA, 1997.
H. Ito, “Pressure Losses in Smooth Pipe Bends,” Journal of
Basic Engineering, March 1960, pp. 131–143.
A. H. Gibson, “On the Flow of Water through Pipes and
­Passages,” Proc. Roy. Soc. London, Ser. A, vol. 83, 1910,
pp. 366–378.
M. Raffel et al., Particle Image Velocimetry: A Practical
Guide, 2d ed., Springer, New York, 2007.
Crane Co., Flow of Fluids through Valves, Fittings, and
Pipe, Crane, Stanford, CT, 2009.
S. A. Berger, L. Talbot, and L.-S. Yao, “Flow in Curved
Pipes,” Annual Review of Fluid Mechanics, vol. 15, 1983,
pp. 461–512.
P. L. Spedding, E. Benard, and G. M. McNally, “Fluid
Flow through 90° Bends,” Developments in Chemical Engineering and Mineral Processing, vol. 12, nos. 1–2, 2004,
pp. 107–128.
References 457
55.
R. R. Kerswell, “Recent Progress in Understanding the
Transition to Turbulence in a Pipe,” Nonlinearity, vol. 18,
2005, pp. R17–R44.
56. B. J. McKeon et al., “Friction Factors for Smooth Pipe
Flow,” J. Fluid Mech., vol. 511, 2004, pp. 41–44.
57. M. Bahrami, M. M. Yovanovich, and J. R. Culham, “Pressure Drop of Fully-Developed Laminar Flow in Micro-
channels of Arbitrary Cross-Section,” J. Fluids
Engineering, vol. 128, Sept. 2006, pp. 1036–1044.
58. G. S. Larraona, A. Rivas, and J. C. Ramos, “Computational
Modeling and Simulation of a Single-Jet Water Meter,” J.
Fluids Engineering, vol. 130, May 2008, pp. 0511021–
05110212.
This chapter is devoted to lift and drag of various bodies immersed in an approaching
stream of fluid. Pictured is the Swiss solar-powered aircraft, Solar Impulse, over the
Golden Gate Bridge. Earlier solar aircraft needed to be towed aloft before flying and were
not able to fly at night. The Solar Impulse is the first solar airplane to fly day and night,
approaching the notion of perpetual flight. The long, high-aspect-ratio wings have more
lift, and less drag, than a short wing of the same area. Its first international flight was
from Switzerland to Brussels, on May 14, 2011. In the summer of 2013, as shown, it flew
from San Francisco to New York City, in five legs. The pilots were Bertrand Piccard and
André Borschberg. [Solar Impulse/Revillard/Rezo]
458
Chapter 7
Flow Past
Immersed Bodies
Motivation. This chapter is devoted to external flows around bodies immersed in
a fluid stream. Such a flow will have viscous (shear and no-slip) effects near the
body surfaces and in its wake, but will typically be nearly inviscid far from the
body. These are unconfined boundary layer flows.
Chapter 6 considered internal flows confined by the walls of a duct. In that
case the viscous boundary layers grow from the sidewalls, meet downstream, and
fill the entire duct. Viscous shear is the dominant effect. For example, the Moody
chart of Fig. 6.13 is essentially a correlation of wall shear stress for long ducts
of constant cross section.
External flows are unconfined, free to expand no matter how thick the
viscous layers grow. Although boundary layer theory (Sec. 7.3) and computational fluid dynamics (CFD) [4] are helpful in understanding external flows,
complex body geometries usually require experimental data on the forces and
moments caused by the flow. Such immersed-body flows are commonly
encountered in engineering ­studies: aerodynamics (airplanes, rockets, projectiles), hydrodynamics (ships, ­submarines, torpedos), transportation (automobiles, trucks, cycles), wind engineering (buildings, bridges, water towers,
wind turbines), and ocean engineering (buoys, breakwaters, pilings, cables,
moored instruments). This chapter provides data and analysis to assist in such
studies.
7.1 Reynolds Number and Geometry Effects
The technique of boundary layer (BL) analysis can be used to compute viscous
effects near solid walls and to “patch” these onto the outer inviscid motion. This
patching is more successful as the body Reynolds number becomes larger, as
shown in Fig. 7.1.
459
460
Chapter 7 Flow Past Immersed Bodies
U
δ≈L
u = 0.99U
Large viscous
displacement
effect
u<U
L
U
ReL = 10
x
Viscous
region
Inviscid region
U
(a)
Small
displacement
effect
δ
L
U
ReL = 107
Fig. 7.1 Comparison of flow past a
sharp flat plate at low and high
Reynolds numbers: (a) laminar,
low-Re flow; (b) high-Re flow.
x
Laminar BL
Turbulent BL
Viscous
Inviscid
region
U
u<U
U
(b)
In Fig. 7.1 a uniform stream U moves parallel to a sharp flat plate of
length L. If the Reynolds number UL/ν is low (Fig. 7.1a), the viscous region
is very broad and extends far ahead and to the sides of the plate. The plate
retards the oncoming stream greatly, and small changes in flow parameters
cause large changes in the pressure distribution along the plate. Thus,
although in principle it should be possible to patch the viscous and inviscid
layers in a mathematical analysis, their interaction is strong and nonlinear [1
to 3]. There is no existing simple theory for external flow analysis at Reynolds numbers from 1 to about 1000. Such thick-shear-layer flows are typically
studied by experiment or by numerical modeling of the flow field on a computer [4].
A high-Reynolds-number flow (Fig. 7.1b) is much more amenable to boundary
layer patching, as first pointed out by Prandtl in 1904. The viscous layers, either
laminar or turbulent, are very thin, thinner even than the drawing shows. We
define the boundary layer thickness δ as the locus of points where the velocity u
parallel to the plate reaches 99 percent of the external velocity U. As we shall
7.1 Reynolds Number and Geometry Effects 461
see in Sec. 7.4, the accepted formulas for flat-plate flow, and their approximate
ranges, are
5.0
δ
Re1/2
≈ µ x
x
0.16
Re1/7
x
10 3 < Rex < 5 × 105
laminar
turbulent 5 × 105 < Rex
(7.1a)
(7.1b)
where Rex = Ux/ν is called the local Reynolds number of the flow along the plate
surface, and x is a distance from the leading edge of the plate. The turbulent flow
formula applies for Rex greater than approximately 5 × 105.
Some computed values from Eq. (7.1) are
Rex
(δ/x)lam
104
105
0.050
0.016
(δ/x)turb
5 × 105
107
108
0.007
0.025
0.016
0.011
The blanks indicate that the formula is not applicable. In all cases these boundary
layers are so thin that their displacement effect on the outer inviscid layer is
negligible. Thus the pressure distribution along the plate can be computed from
inviscid theory as if the boundary layer were not even there. This external pressure field then “drives” the boundary layer flow, acting as a forcing function in
the momentum equation along the surface. We shall explain this boundary layer
theory in Secs. 7.4 and 7.5.
External flow can be classified based on the body shapes. Bodies can be sorted
into two general categories, depending on whether they are slender bodies, such
as thin plates and airfoils parallel to the oncoming stream, or blunt bodies, such
as cylinders and spheres. Different approaches to the slender and blunt body have
to be used in fluid mechanics, and we shall discuss them in separate sections in
this chapter.
For slender bodies, we conclude that this assumption of negligible interaction
between the boundary layer and the outer pressure distribution is an excellent
approximation. The boundary layer theory simplifies the analysis of high Reynolds number flows, thereby allowing solution to this category of external flow
problems.
For a blunt-body flow described in Sec. 7.6, however, even at very high Reynolds numbers, there is a discrepancy in the viscous–inviscid patching concept.
Figure 7.2 shows two sketches of flow past a two- or three-dimensional blunt
body. In the ideal­ized sketch (7.2a), there is a thin film of boundary layer about
the body and a ­narrow sheet of viscous wake in the rear. The patching theory
would be glorious for this picture, but it is false. In the actual flow (Fig. 7.2b),
the boundary layer is thin on the front, or windward, side of the body, where the
pressure decreases along the surface (favorable pressure gradient). But in the rear
the boundary layer encounters increasing pressure (adverse pressure gradient) and
breaks off, or separates, into a broad, pulsating wake. (See Fig. 5.2a for a
462
Chapter 7 Flow Past Immersed Bodies
Beautifully behaved
but mythically thin
boundary layer
and wake
Red = 105
Thin front
boundary layer
Outer stream grossly
perturbed by broad flow
separation and wake
(a)
Red = 105
Fig. 7.2 Illustration of the strong
­interaction between viscous and
­inviscid regions in the rear of
blunt-body flow: (a) idealized
and definitely false picture of
blunt-body flow; (b) actual picture
of blunt-body flow.
(b)
photograph of a specific example.) The mainstream is deflected by this wake, so
that the external flow is quite different from the prediction from inviscid theory
with the addition of a thin boundary layer.
The theory of strong interaction between blunt-body viscous and inviscid layers
is not well developed. Flows like that of Fig. 7.2b are normally studied experimentally or with CFD [4]. Reference 5 is an example of efforts to improve the theory
of separated flows. Reference 6 is another textbook devoted to separated flow.
EXAMPLE 7.1
A long, thin, flat plate is placed parallel to a 20-ft/s stream of water at 68°F. At what distance
x from the leading edge will the boundary layer thickness be 1 in?
Solution
∙
∙
∙
∙
Assumptions: Flat-plate flow, with Eqs. (7.1) applying in their appropriate ranges.
Approach: Guess laminar flow first. If contradictory, try turbulent flow.
Property values: From Table A.1 for water at 68°F, ν ≈ 1.082 E−5 ft2/s.
Solution step 1: With δ = 1 in = 1/12 ft, try laminar flow, Eq. (7.1a):
δ
x
∣
lam
=
5
(Ux/ν) 1/2
or
1/12 ft
5
=
x
[ (20 ft/s)x/(1.082 E−5 ft2/s) ] 1/2
Solve for
x ≈ 513 ft
7.2 Momentum Integral Estimates 463
Pretty long plate! This does not sound right. Check the local Reynolds number:
Rex =
Ux (20 ft/s) (513 ft)
=
= 9.5 E8
ν
1.082 E−5 ft2/s
(!)
This is impossible, since laminar boundary layer flow only persists up to about 5 × 105
(or, with special care to avoid disturbances, up to 3 × 106).
∙ Solution step 2: Try turbulent flow, Eq. (7.1b):
δ
0.16
=
x (Ux/ν) 1/7
or
1/12 ft
0.16
=
x
[ (20 ft/s)x/(1.082 E−5 ft2/s) ] 1/7
Solve for x ≈ 5.17 ft
Ans.
Check Rex = (20 ft/s)(5.17 ft)/(1.082 E−5 ft2/s) = 9.6 E6 > 5 × 105. OK, turbulent
flow.
∙ Comments: The flow is turbulent, and the inherent ambiguity of the theory is
resolved.
7.2 Momentum Integral Estimates
When we derived the momentum integral relation, Eq. (3.37), and applied it to
a flat-plate boundary layer in Example 3.9, we promised to consider it further in
Chap. 7. Well, here we are! Let us review the problem, using Fig. 7.3.
A shear layer of unknown thickness grows along the sharp flat plate in Fig. 7.3.
The no-slip wall condition retards the flow, making it into a rounded profile
u(x, y), which merges into the external velocity U = constant at a “thickness”
y = δ(x). By utilizing the control volume of Fig. 3.11, we found (without making
any assumptions about laminar versus turbulent flow) in Example 3.9 that the
drag force on the plate is given by the following momentum integral across the
exit plane:
D(x) = ρb
∫
δ(x)
0
u(U − u) dy
(7.2)
where b is the plate width into the paper and the integration is carried out along
a vertical plane x = constant. You should review the momentum integral relation
(3.37) and its use in Example 3.9.
y
U
U
p = pa
δ (x)
τ w(x)
Fig. 7.3 Growth of a boundary
layer on a flat plate. The thickness
is exaggerated.
u(x, y)
x
x=0
x=L
464
Chapter 7 Flow Past Immersed Bodies
Kármán’s Analysis of the Flat Plate
Equation (7.2) was derived in 1921 by Kármán [7], who wrote it in the convenient
form of the momentum thickness θ:
D(x) = ρbU2θ
θ=
∫
δ
0
u
u
1 − ) dy
U(
U
(7.3)
Momentum thickness is thus a measure of total plate drag. Kármán then noted
that the drag also equals the integrated wall shear stress along the plate:
∫
x
D(x) = b τw (x) dx
0
dD
= bτw
dx
or
(7.4)
Meanwhile, the derivative of Eq. (7.3), with U = constant, is
dD
dθ
= ρbU2
dx
dx
By comparing this with Eq. (7.4), Kármán arrived at what is now called the
momentum integral relation for flat-plate boundary layer flow:
τw = ρU2
dθ
dx
(7.5)
It is valid for either laminar or turbulent flat-plate flow.
To get a numerical result for laminar flow, Kármán assumed that the velocity
profiles had an approximately parabolic shape
2y y2
u(x, y) ≈ U ( − 2 )
δ
δ
0 ≤ y ≤ δ(x)
(7.6)
which makes it possible to estimate both momentum thickness and wall shear:
θ=
2y y2
2y y2
2
−
1
−
+ 2 ) dy ≈ δ
2) (
(δ
δ
15
δ
δ
0
∫
δ
τw = μ
∂u
∂y
∣
y=0
≈
2μU
δ
By substituting (7.7) into (7.5) and rearranging, we obtain
ν
δ dδ ≈ 15 dx
U
(7.7)
(7.8)
where ν = µ/ρ. We can integrate from 0 to x, assuming that δ = 0 at x = 0, the
leading edge:
or
1 2 15νx
δ =
2
U
δ
ν 1/2
5.5
≈ 5.5
= 1/2 (
)
x
Ux
Rex
(7.9)
7.2 Momentum Integral Estimates 465
This is the desired thickness estimate. It is all approximate, of course, part of
Kármán’s momentum integral theory [7], but it is startlingly accurate, being only
10 percent higher than the known accepted solution for laminar flat-plate flow,
which we gave as Eq. (7.1a).
By combining Eqs. (7.9) and (7.7), we also obtain a shear stress estimate along
the plate:
8
1/2
2τw
0.73
15
cf =
≈(
= 1/2 (7.10)
2
)
Re
ρU
Rex
x
Again this estimate, in spite of the crudeness of the profile assumption [Eq. (7.6)]
is only 10 percent higher than the known exact laminar-plate-flow solution cf =
0.664/Re1/2
x , treated in Sec. 7.4. The dimensionless quantity cf, called the skin friction
coefficient, is analogous to the friction factor f in ducts.
A boundary layer can be judged as “thin” if, say, the ratio δ/x is less than
about 0.1. This occurs at δ/x = 0.1 = 5.0/Re1/2
x or at Rex = 2500. For Rex less
than 2500, we can estimate that boundary layer theory fails because the thick
layer has a ­significant effect on the outer inviscid flow. The upper limit on Rex
for laminar flow is about 3 × 106, where measurements on a smooth flat plate
[8] show that the flow undergoes transition to a turbulent boundary layer. From
3 × 106 upward the turbulent Reynolds number may be arbitrarily large, and a
practical limit at present is 5 × 1010 for oil supertankers.
Displacement Thickness
Another interesting effect of a boundary layer is its small but finite displacement
of the outer streamlines. As shown in Fig. 7.4, outer streamlines must deflect outward a distance δ*(x) to satisfy conservation of mass between the inlet and outlet:
h
δ
0
0
∫ ρUb dy = ∫ ρub dy
δ = h + δ*
(7.11)
The quantity δ* is called the displacement thickness of the boundary layer. To
relate it to u(y), cancel ρ and b from Eq. (7.11), evaluate the left integral, and
slyly add and subtract U from the right integrand:
Uh =
∫
δ
0
δ
(U + u − U) dy = U(h + δ*) +
∫ (1 − Uu ) dy
∫ (u − U) dy
0
δ
or
δ* =
y = h + δ*
y
U
Fig. 7.4 Displacement effect of a
boundary layer.
U
U
y=h
h
Outer streamline
h
u
0
x
(7.12)
0
Simulated
effect
δ*
466
Chapter 7 Flow Past Immersed Bodies
Thus the ratio of δ*/δ varies only with the dimensionless velocity profile
shape u/U.
Introducing our profile approximation (7.6) into (7.12), we obtain by integration this approximate result:
δ* ≈
1
δ
3
δ* 1.83
≈ 1/2 x
Rex
(7.13)
These estimates are only 6 percent away from the exact solutions for laminar flatplate flow given in Sec. 7.4: δ* = 0.344δ = 1.721x/Re1/2
x . Since δ* is much smaller
than x for large Rex and the outer streamline slope V/U is proportional to δ*, we
conclude that the velocity normal to the wall is much smaller than the velocity
parallel to the wall. This is a key assumption in boundary layer theory (Sec. 7.3).
We also conclude from the success of these simple parabolic estimates that
Kármán’s momentum integral theory is effective and useful. Many details of this
theory are given in Refs. 1 to 3.
EXAMPLE 7.2
Are low-speed, small-scale air and water boundary layers really thin? Consider flow at
U = 1 ft/s past a flat plate 1 ft long. Compute the boundary layer thickness at the
trailing edge for (a) air and (b) water at 68°F.
Solution
Part (a)
From Table A.2, νair ≈ 1.61 E−4 ft2/s. The trailing-edge Reynolds number thus is
ReL =
(1 ft/s) (1 ft)
UL
=
= 6200
ν
1.61 E−4 ft2/s
Since this is less than 5 × 105, the flow is presumed laminar, and since it is greater
than 2500, the boundary layer is reasonably thin. From Eq. (7.1a), the predicted laminar
thickness is
δ
=
x
or, at x = 1 ft,
5.0
√6200
= 0.0634
δ = 0.0634 ft ≈ 0.76 in
Ans. (a)
Part (b)
From Table A.1, νwater ≈ 1.08 E−5 ft2/s. The trailing-edge Reynolds number is
(1 ft/s) (1 ft)
≈ 92,600
1.08 E−5 ft2/s
This again satisfies the laminar and thinness conditions. The boundary layer thickness is
ReL =
δ
≈
x
5.0
√92,600
= 0.0164
7.3 The Boundary Layer Equations 467
or, at x = 1 ft,
δ = 0.0164 ft ≈ 0.20 in
Ans. (b)
Thus, even at such low velocities and short lengths, both airflows and water flows
satisfy the boundary layer approximations.
7.3 The Boundary Layer Equations
In Chaps. 4 and 6, we learned that there are several dozen known analytical
laminar flow solutions [1 to 3]. None are for external flow around immersed
bodies, although this is one of the primary applications of fluid mechanics. No
exact solutions are known for turbulent flow, whose analysis typically uses
­empirical modeling laws to relate time-mean variables.
There are presently three techniques used to study external flows: (1) numerical (computer) solutions, (2) experimentation, and (3) boundary layer theory.
Computational fluid dynamics is now well developed and described in advanced
texts such as that by Anderson [4]. Thousands of computer solutions and models
have been published; execution times, mesh sizes, and graphical presentations are
improving each year. Both laminar and turbulent flow solutions have been published, and turbulence modeling is a current research topic [9]. However, the topic
of CFD is beyond our scope here.
Experimentation is the most common method of studying external flows.
­Chapter 5 outlined the technique of dimensional analysis, and we shall give
many nondimensional experimental data for flow over immersed blunt bodies
in Sec. 7.6.
The third tool is boundary layer theory, first formulated by Ludwig Prandtl in
1904. We shall follow Prandtl’s ideas here and make certain order-of-magnitude
assumptions to greatly simplify the Navier–Stokes equations (4.38) into boundary
layer equations that are solved relatively easily and patched onto the outer inviscid flow field.
One of the great achievements of boundary layer theory is its ability to predict
the flow separation that occurs in adverse (positive) pressure gradients, as illustrated in Fig. 7.2b. Before 1904, when Prandtl published his pioneering paper, no
one realized that such thin shear layers could cause such a gross effect as flow
separation. Even today, however, boundary layer theory cannot accurately predict
the behavior of the separated-flow region and its interaction with the outer flow.
Modern research [4, 9] has focused on detailed CFD simulations of separated
flow, and the resultant wakes, to gain further insight.
Derivation for Two-Dimensional Flow
We consider only steady two-dimensional incompressible viscous flow with the
x direction along the wall and y normal to the wall, as in Fig. 7.3.1 We neglect
gravity, which is important only in boundary layers where fluid buoyancy is
1
For a curved wall, x can represent the arc length along the wall and y can be everywhere normal to x with negligible change in the boundary layer equations as long as the radius of curvature
of the wall is large compared with the boundary layer thickness [1 to 3].
468
Chapter 7 Flow Past Immersed Bodies
dominant [2, sec. 4.14]. From Chap. 4, the complete equations of motion consist
of continuity and the x- and y-momentum relations:
∂u ∂υ
+
= 0
∂x ∂y
ρ (u
ρ (u
(7.14a)
∂p
∂u
∂u
∂ 2u ∂ 2u
+ υ ) = − + μ ( 2 + 2 )
∂x
∂y
∂x
∂x
∂y
∂p
∂υ
∂υ
∂ 2υ ∂ 2υ
+ υ ) = − + μ ( 2 + 2 )
∂x
∂y
∂y
∂x
∂y
(7.14b)
(7.14c)
These should be solved for u, υ, and p subject to typical no-slip, inlet, and exit
boundary conditions, but in fact they are too difficult to handle for most external
flows except with CFD.
In 1904, Prandtl correctly deduced that a shear layer must be very thin if the
Reynolds number is large, so that the following approximations apply:
υ ≪ u
Velocities:
∂u
∂u
≪
∂x
∂y
Rates of change:
Reynolds number:
Rex =
(7.15a)
∂υ
∂υ
≪
∂x
∂y
(7.15b)
Ux
≫ 1
ν
(7.15c)
Our discussion of displacement thickness in the previous section was intended to
justify these assumptions.
Applying these approximations to Eq. (7.14c) results in a powerful simplification:
ρ (u
∂p
∂υ
∂υ
∂ 2υ
∂ 2υ
+
ρ
υ
=
−
+
μ
+
μ
2
( ∂y )
( ∂x )
( ∂y2 )
∂x )
∂y
small
small
very small
∂p
≈ 0 or p ≈ p(x)
∂y
only
small
(7.16)
In other words, the y-momentum equation can be neglected entirely, and the pressure varies only along the boundary layer, not through it. The pressure gradient
term in Eq. (7.14b) is assumed to be known in advance from Bernoulli’s equation
applied to the outer inviscid flow:
∂p dp
dU
=
= −ρU
∂x dx
dx
(7.17)
Presumably we have already made the inviscid analysis and know the distribution
of U(x) along the wall (Chap. 8).
Meanwhile, one term in Eq. (7.14b) is negligible due to Eqs. (7.15):
∂ 2u
∂ 2u
≪
2
∂x
∂y2
(7.18)
However, neither term in the continuity relation (7.14a) can be neglected—another
warning that continuity is always a vital part of any fluid flow analysis.
7.4 The Flat-Plate Boundary Layer 469
The net result is that the three full equations of motion (7.14) are reduced to
Prandtl’s two boundary layer equations for two-dimensional incompressible flow:
∂u ∂υ
+
= 0
∂x ∂y
Continuity:
Momentum along wall:
u
∂u
∂u
dU 1 ∂τ
+υ
≈U
+
ρ ∂y
∂x
∂y
dx
∂u
∂y
τ= µ
∂u
μ
− ρu′υ′
∂y
μ
where
(7.19a)
(7.19b)
laminar flow
turbulent flow
These are to be solved for u(x, y) and υ(x, y), with U(x) assumed to be a known
function from the outer inviscid flow analysis. There are two boundary conditions
on u and one on υ:
At y = 0 (wall):
u=υ=0
(no slip)
(7.20a)
As y = δ(x) (other stream):
u = U(x)
(patching)
(7.20b)
Unlike the Navier–Stokes equations (7.14), which are mathematically elliptic and
must be solved simultaneously over the entire flow field, the boundary layer equations (7.19) are mathematically parabolic and are solved by beginning at the
leading edge and marching downstream as far as you like, stopping at the separation point or earlier if you prefer.2
The boundary layer equations have been solved for scores of interesting cases
of internal and external flow for both laminar and turbulent flow, utilizing the
inviscid distribution U(x) appropriate to each flow. Full details of boundary layer
theory and results and comparison with experiment are given in Refs. 1 to 3.
Here we shall confine ourselves primarily to flat-plate solutions (Sec. 7.4).
7.4 The Flat-Plate Boundary Layer
The classic and most often used solution of boundary layer theory is for flat-plate
flow, as in Fig. 7.3, which can represent either laminar or turbulent flow.
Laminar Flow
For laminar flow past the plate, the boundary layer equations (7.19) can be solved
exactly for u and υ, assuming that the free-stream velocity U is constant (dU/dx
= 0). The solution was given by Prandtl’s student Blasius, in his 1908 dissertation
from Göttingen. With an ingenious coordinate transformation, Blasius showed that
the dimensionless velocity profile u/U is a function only of the single composite
dimensionless variable (y)[U/(νx)]1/2:
u
= f ′(η)
U
2
U 1/2
η = y( ) νx
For further mathematical details, see Ref. 2, Sec. 2.8.
(7.21)
470
Chapter 7 Flow Past Immersed Bodies
Table 7.1 The Blasius Velocity
Profile [1 to 3]
y[U/(νx)]1/2
u/U
y[U/(νx)]1/2
u/U
0.0
0.2
0.4
0.6
0.8
1.0
1.2
1.4
1.6
1.8
2.0
2.2
2.4
2.6
0.0
0.06641
0.13277
0.19894
0.26471
0.32979
0.39378
0.45627
0.51676
0.57477
0.62977
0.68132
0.72899
0.77246
2.8
3.0
3.2
3.4
3.6
3.8
4.0
4.2
4.4
4.6
4.8
5.0
q
0.81152
0.84605
0.87609
0.90177
0.92333
0.94112
0.95552
0.96696
0.97587
0.98269
0.98779
0.99155
1.00000
where the prime denotes differentiation with respect to η. Substitution of (7.21)
into the boundary layer equations (7.19) reduces the problem, after much algebra,
to a single third-order nonlinear ordinary differential equation for f [1–3]:
f ‴ + 12 ff ″ = 0
(7.22)
The boundary conditions (7.20) become
At y = 0:
As y → ∞:
f(0) = f ′(0) = 0
(7.23a)
f ′(∞ ) → 1.0
(7.23b)
This is the Blasius equation, for which accurate solutions have been obtained only
by numerical integration. Some tabulated values of the velocity profile shape f'(η)
= u/U are given in Table 7.1.
Since u/U approaches 1.0 only as y → ∞, it is customary to select the
boundary layer thickness δ as that point where u/U = 0.99. From the table,
this occurs at η ≈ 5.0:
δ99%(
or
δ
5.0
≈
x Re1/2
x
U 1/2
≈ 5.0
νx )
Blasius (1908)
(7.24)
With the profile known, Blasius, of course, could also compute the wall shear
and displacement thickness:
cf =
0.664
Re1/2
x
δ* 1.721
=
x
Re1/2
x
(7.25)
Notice how close these are to our integral estimates, Eqs. (7.9), (7.10), and (7.13).
When cf is converted to dimensional form, we have
τw (x) =
0.332ρ1/2μ1/2U1.5
x1/2
7.4 The Flat-Plate Boundary Layer 471
The wall shear drops off with x1/2 because of boundary layer growth and varies
as velocity to the 1.5 power. This is in contrast to fully developed laminar pipe
flow, where τw ∝ U and is independent of x.
If τw(x) is substituted into Eq. (7.4), we compute the total drag force:
∫
x
D(x) = b τw (x) dx = 0.664bρ1/2μ1/2U1.5x1/2
0
(7.26)
The drag increases only as the square root of the plate length. The nondimensional
drag coefficient is defined as
CD =
2D(L)
2
ρU bL
=
1.328
= 2cf (L) Re1/2
L
(7.27)
Thus, for laminar plate flow, CD equals twice the value of the skin friction coefficient at the trailing edge. This is the drag on one side of the plate.
Kármán pointed out that the drag could also be computed from the momentum
relation (7.2). In dimensionless form, Eq. (7.2) becomes
CD =
2
L
∫
δ
0
u
u
1 − ) dy
U(
U
(7.28)
This can be rewritten in terms of the momentum thickness at the trailing edge:
CD =
2θ(L)
L
(7.29)
Computation of θ from the profile u/U or from CD gives
θ 0.664
=
x
Re1/2
x
laminar flat plate
(7.30)
Since δ is so ill defined, the momentum thickness, being definite, is often used
to correlate data taken for a variety of boundary layers under differing conditions.
The ratio of displacement to momentum thickness, called the dimensionless-profile shape factor, is also useful in integral theories. For laminar flat-plate flow
H=
δ* 1.721
=
= 2.59
θ
0.664
(7.31)
A large shape factor then implies that boundary layer separation is about to occur.
If we plot the Blasius velocity profile from Table 7.1 in the form of u/U versus y/δ, we can see why the simple integral theory guess, Eq. (7.6), was such a
great success. This is done in Fig. 7.5. The simple parabolic approximation is not
far from the true Blasius profile; hence its momentum thickness is within 10
percent of the true value. Also shown in Fig. 7.5 are three typical turbulent flatplate velocity profiles. Notice how strikingly different in shape they are from the
laminar profiles. Instead of decreasing parabolically to zero, the turbulent profiles
are very flat and then drop off sharply at the wall. As you might guess, they
follow the logarithmic law shape and thus can be analyzed by momentum integral
theory if this shape is properly represented.
472
Chapter 7 Flow Past Immersed Bodies
1.0
Turbulent
0.8
Seventh
root profile,
Eq. (7.39)
105 = Rex
106
107
0.6
u
U
0.4
Exact Blasius profile
for all laminar Rex
(Table 7.1)
0.2
Fig. 7.5 Comparison of
dimensionless laminar and turbulent
flat-plate velocity profiles.
0
Parabolic
approximation,
Eq. (7.6)
0.2
0.4
0.6
0.8
1.0
y
δ
Transition to Turbulence
The laminar flat-plate boundary layer eventually becomes turbulent. This will
occur for any stream velocity and any fluid, provided the plate is long enough.
This is because the parameter that governs the transition of flow from laminar
to turbulent flow is the local Reynolds number, Rex = Ux/v, where x is the
distance from the leading edge of the plate. The value of the transition Reynolds
number is a rather complex function of various parameters, including the surface roughness, disturbances from the flow outside the boundary layer, the
curvature of the surface, etc. With care in polishing the plate surface and keeping the free stream quiet, one can delay the transition Reynolds number to Rex,tr
≈ 3 E6 [8]. However, for typical commercial surfaces and gusty free streams,
a more realistic value is
Rex,tr ≈ 5 E5.
It should be noted that the critical Reynolds number for transition used for external flows is quite different from that for internal flows in Chap. 6, where Red,crit
≈ 2300 in Eq. (6.2) is based on the diameter of the pipe. Therefore, if the Reynolds number for a pipe flow is less than 2300, the flow is always laminar no
matter how long the pipe is.
Transition from laminar to turbulent flow involves instability of the flow field,
similar to what we described in Sec. 6.1. One consequence of the transition is a
noticeable change in the shape of the boundary layer velocity profile illustrated
in Fig. 7.5.
7.4 The Flat-Plate Boundary Layer 473
EXAMPLE 7.3
A sharp flat plate with L = 50 cm and b = 3 m is parallel to a stream of velocity
2.5 m/s. Find the drag on one side of the plate, and the boundary thickness δ at the
trailing edge, for (a) air and (b) water at 20°C and 1 atm.
Solution
∙ Assumptions: Laminar flat-plate flow, but we should check the Reynolds numbers.
∙ Approach: Find the Reynolds number and use the appropriate boundary layer formulas.
∙ Property values: From Table A.2 for air at 20°C, ρ = 1.2 kg/m3, ν = 1.5 E−5 m2/s.
From Table A.1 for water at 20°C, ρ = 998 kg/m3, ν = 1.005 E−6 m2/s.
∙ (a) Solution for air: Calculate the Reynolds number at the trailing edge:
ReL =
UL (2.5 m/s) (0.5 m)
=
= 83,300 < 5 E5 therefore assuredly laminar
νair
1.5 E−5 m2/s
The appropriate thickness relation is Eq. (7.24):
δ
5
5
=
=
= 0.0173, or δx=L = 0.0173(0.5 m) ≅ 0.0087 m
L Re1/2
(83,300) 1/2
L
Ans. (a)
The laminar boundary layer is only 8.7 mm thick. The drag coefficient follows from
Eq. (7.27):
CD =
1.328
1.328
=
= 0.0046
1/2
Re1/2
(83,300)
L
1.2 kg/m3
ρ
or Done side = CD U2bL = (0.0046)
(2.5 m/s) 2 (3 m) (0.5 m) ≈ 0.026 N Ans. (a)
2
2
∙ Comment (a): This is purely friction drag and is very small for gases at low velocities.
∙ (b) Solution for water: Again calculate the Reynolds number at the trailing edge:
ReL =
(2.5 m/s) (0.5 m)
UL
=
= 1.24 E6 > 5 E5 therefore it might be turbulent
νwater
1.005 E−6 m2/s
This is a quandary. If the plate is rough or encounters disturbances, the flow at the
trailing edge will be turbulent. Let us assume a smooth, undisturbed plate, which will
remain laminar. Then again the appropriate thickness relation is Eq. (7.24):
δ
5
5
=
=
= 0.00448 or
L Re1/2
(1.24 E6) 1/2
L
δx=L = 0.00448(0.5 m) ≅ 0.0022 m Ans. (b)
This is four times thinner than the air result in part (a), due to the high laminar
Reynolds number. Again the drag coefficient follows from Eq. (7.27):
CD =
1.328
1.328
=
= 0.0012
1/2
ReL
(1.24 E6) 1/2
998 kg/m3
ρ
or Done side = CD U2bL = (0.0012)
(2.5 m/s) 2 (3 m) (0.5 m) ≈ 5.6 N Ans. (b)
2
2
∙ Comment (b): The drag is 215 times larger for water, although CD is lower, reflecting that water is 56 times more viscous and 830 times denser than air. From Eq.
474
Chapter 7 Flow Past Immersed Bodies
(7.26), for the same U and x, the water drag should be (56)1/2(830)1/2 ≈ 215 times
higher. Note: If transition to turbulence had occurred at Rex = 5 E5 (at about x = 20
cm), the drag would be about 2.5 times higher, and the trailing edge thickness about
four times higher than for fully laminar flow.
Turbulent Flow
There is no exact theory for turbulent flat-plate flow, although there are many elegant
computer solutions of the boundary layer equations using various empirical models
for the turbulent eddy viscosity [9]. The most widely accepted result is simply an
integral analysis similar to our study of the laminar profile approximation (7.6).
We begin with Eq. (7.5), which is valid for laminar or turbulent flow. We write
it here for convenient reference:
τw (x) = ρU2
dθ
dx
(7.32)
From the definition of cf, Eq. (7.10), this can be rewritten as
cf = 2
dθ
dx
(7.33)
Now recall from Fig. 7.5 that the turbulent profiles are nowhere near parabolic.
Going back to Fig. 6.10, we see that flat-plate flow is very nearly logarithmic,
with a slight outer wake and a thin viscous sublayer. Therefore, just as in turbulent pipe flow, we assume that the logarithmic law (6.28) holds all the way across
the boundary layer
τw 1/2
u
1 yu*
≈ ln
+ B u* = ( ) ν
ρ
u* κ
(7.34)
with, as usual, κ = 0.41 and B = 5.0. At the outer edge of the boundary layer,
y = δ and u = U, and Eq. (7.34) becomes
U
1 δu*
= ln
+ B
ν
u* κ
(7.35)
But the definition of the skin friction coefficient, Eq. (7.10), is such that the
f­ollowing identities hold:
cf 1/2
U
2 1/2 δu*
≡ Reδ
≡( )
(7.36)
(2) cf
ν
u*
Therefore, Eq. (7.35) is a skin friction law for turbulent flat-plate flow:
cf 1/2
2 1/2
≈
2.44
ln
Re
δ
( cf )
[
( 2 ) ] + 5.0
(7.37)
It is a complicated law, but we can at least solve for a few values and list them:
Reδ
cf
104
105
106
107
0.00493
0.00315
0.00217
0.00158
7.4 The Flat-Plate Boundary Layer 475
Following a suggestion of Prandtl, we can forget the complex log friction law
(7.37) and simply fit the numbers in the table to a power-law approximation:
cf ≈ 0.02 Reδ−1/6
(7.38)
This we shall use as the left-hand side of Eq. (7.33). For the right-hand side, we
need an estimate for θ(x) in terms of δ(x). If we use the logarithmic law profile
(7.34), we shall be up to our hips in logarithmic integrations for the momentum
thickness. Instead we follow another suggestion of Prandtl, who pointed out that
the turbulent profiles in Fig. 7.5 can be approximated by a one-seventh-power
law:
y 1/7
u
≈
( U )turb ( δ ) (7.39)
This is shown as a dashed line in Fig. 7.5. It is an excellent fit to the lowReynolds-number turbulent data, which were all that were available to Prandtl at
the time. With this simple approximation, the momentum thickness (7.28) can
easily be evaluated:
θ≈
∫
δ
0
y 1/7
y 1/7
7
1
−
(δ) [
( δ ) ]dy = 72 δ
(7.40)
We accept this result and substitute Eqs. (7.38) and (7.40) into Kármán’s momentum law (7.33):
cf = 0.02 Re−1/6
=2
δ
or
Re−1/6
= 9.72
δ
d 7
δ
dx ( 72 )
d(Reδ )
dδ
= 9.72
dx
d(Rex )
(7.41)
Separate the variables and integrate, assuming δ = 0 at x = 0:
Reδ ≈ 0.16 Re6/7
x
or
δ 0.16
≈
x Re1/7
x
(7.42)
Thus the thickness of a turbulent boundary layer increases as x6/7, far more rapidly
than the laminar increase x1/2. Equation (7.42) is the solution to the problem,
because all other parameters are now available. For example, combining Eqs.
(7.42) and (7.38), we obtain the friction variation
cf ≈
0.027
Re1/7
x
(7.43)
Writing this out in dimensional form, we have
τw,turb ≈
0.0135μ1/7ρ6/7U13/7
(7.44)
x1/7
Turbulent plate friction drops slowly with x, increases nearly as ρ and U2, and is
rather insensitive to viscosity.
476
Chapter 7 Flow Past Immersed Bodies
We can evaluate the drag coefficient by integrating the wall friction:
L
D=
or
CD =
∫ τ b dx
w
0
2D
=
ρU2 bL
CD =
∫ c d (Lx )
1
f
0
0.031 7
= cf (L) 6
Re1/7
L
(7.45)
Then CD is only 16 percent greater than the trailing-edge skin friction coefficient
[compare with Eq. (7.27) for laminar flow].
The displacement thickness can be estimated from the one-seventh-power law
and Eq. (7.12):
∫ [1 − (δy)
δ
δ* ≈
1/7
0
1
]dy = 8 δ
(7.46)
The turbulent flat-plate shape factor is approximately
H=
δ*
=
θ
1
8
7
72
= 1.3
(7.47)
These are the basic results of turbulent flat-plate theory.
Figure 7.6 shows flat-plate drag coefficients for both laminar and turbulent
flow conditions. The smooth-wall relations (7.27) and (7.45) are shown, along
with the effect of wall roughness, which is quite strong. The proper roughness
parameter here is x/ε or L/ε, by analogy with the pipe parameter ε/d. In the fully
rough regime, CD is independent of the Reynolds number, so that the drag varies
exactly as U2 and is independent of µ. Reference 2 presents a theory of rough
flat-plate flow, and Ref. 1 gives a curve fit for skin friction and drag in the fully
rough regime:
x −2.5
cf ≈ (2.87 + 1.58 log ) ε
(7.48a)
L −2.5
CD ≈ (1.89 + 1.62 log ) ε
(7.48b)
Equation (7.48b) is plotted to the right of the dashed line in Fig. 7.6. The figure
also shows the behavior of the drag coefficient in the transition region 5 × 105 <
ReL < 8 × 107, where the laminar drag at the leading edge is an appreciable fraction
of the total drag. Schlichting [1] suggests the following curve fits for these transition
drag curves, depending on the Reynolds number Retrans where transition begins:
0.031 1440
−
ReL
Re1/7
L
CD ≈ µ
0.031 8700
−
ReL
Re1/7
L
Retrans = 5 × 105
Retrans = 3 × 106
(7.49a)
(7.49b)
7.4 The Flat-Plate Boundary Layer 477
0.014
200
Fully rough
Eq. (7.48b)
0.012
L
Ɛ = 300
500
0.010
1000
0.008
2000
CD
0.006
5000
104
2 × 104
0.004
2 × 105
106
Transition
0.002
Fig. 7.6 Drag coefficient of laminar
and turbulent boundary layers on
smooth and rough flat plates. This
chart is the flat-plate analog of the
Moody diagram of Fig. 6.13.
5 × 104
Turbulent
smooth
Eq. (7.45)
Laminar:
Eq. (7.27)
0
105
Eq. (7.49)
106
107
108
109
ReL
For laminar flow over the entire plate (xc/L > 1), Eqs. (7.24), (7.25), and (7.27)
may be used to compute the BL thickness, local friction, and drag coefficients.
If transition occurs at a point very close to the leading edge (xc/L ≪ 1), the
boundary layer can be approximated as turbulent throughout, so that Eqs. (7.42),
(7.43), and (7.45) may be adopted. However, when transition occurs around the
center of the plate (xc/L ≈ 1/2), the surface drag coefficients will be influenced
by conditions in both the laminar and the turbulent boundary layers, and
Eq. (7.49) may be used for this situation.
EXAMPLE 7.4
A hydrofoil 1.2 ft long and 6 ft wide is placed in a seawater flow of 40 ft/s, with ρ =
1.99 slugs/ft3 and ν = 0.000011 ft2/s. (a) Estimate the boundary layer thickness at the
end of the plate. Estimate the friction drag for (b) turbulent smooth-wall flow from the
leading edge, (c) laminar turbulent flow with Retrans = 5 × 105, and (d) turbulent roughwall flow with ε = 0.0004 ft.
478
Chapter 7 Flow Past Immersed Bodies
Solution
Part (a)
The Reynolds number is
UL (40 ft/s) (1.2 ft)
=
= 4.36 × 106
ν
0.000011 ft2/s
Thus the trailing-edge flow is certainly turbulent. The maximum boundary layer
­thickness would occur for turbulent flow starting at the leading edge. From Eq. (7.42),
ReL =
δ(L)
0.16
=
= 0.018
L
(4.36 × 106 ) 1/7
or
δ = 0.018(1.2 ft) = 0.0216 ft
Ans. (a)
This is 7.5 times thicker than a fully laminar boundary layer at the same Reynolds
number.
Part (b)
For fully turbulent smooth-wall flow, the drag coefficient on one side of the plate is, from
Eq. (7.45),
CD =
0.031
= 0.00349
(4.36 × 106 ) 1/7
Then the drag on both sides of the foil is approximately
D = 2CD ( 12 ρU2 )bL = 2(0.00349) ( 12 ) (1.99) (40) 2 (6.0) (1.2) = 80 lbf
Ans. (b)
Part (c)
With a laminar leading edge and Retrans = 5 × 105, Eq. (7.49a) applies:
CD = 0.00349 −
1440
= 0.00316
4.36 × 106
The drag can be recomputed for this lower drag coefficient:
D = 2CD ( 12 ρU2 )bL = 72 lbf
Ans. (c)
Part (d)
Finally, for the rough wall, we calculate
L
1.2 ft
=
= 3000
ε 0.0004 ft
From Fig. 7.6 at ReL = 4.36 × 106, this condition is just inside the fully rough regime.
Equation (7.48b) applies:
CD = (1.89 + 1.62 log 3000) −2.5 = 0.00644
and the drag estima